Synchronized Regenerators and an Improved Bland/Ewing Thermochemical Cycle

ABSTRACT

For efficiently exchanging heat between two streams of fluid at approximately equal pressure while simultaneously reducing the internal volume and general overall mass of the heat exchange means per quantity of heat exchanged over time, a means termed a Synchronized Thermal Regenerator Exchange Pump (STREP) is proposed.

CROSS-REFERENCE TO RELATED APPLICATIONS

This application claims priority to and the benefit of U.S. ProvisionalPatent Application No. 63/393,960, filed Jul. 31, 2022, and to U.S.Provisional Patent Application No. 63/439,781, filed Jan. 18, 2023, theentire content of each of which is incorporated herein by reference.

REFERENCES

This field is related in part to the invention disclosed in U.S.Provisional Patent Application No. 63/393,960 termed a SynchronizedThermal Regenerator Exchange Pump (STREP), in part to a heat enginecycle invention disclosed in U.S. Provisional Patent Application No.63/439,781 termed a Synchronizing Displacer (SD) Valved Cell (SD-VC)engine, and in part on U.S. Pat. Nos. 3,225,538, 3,067,594, 3,871,179,4,817,388, and 5,215,691.

Reference is also made to U.S. patent application Ser. Nos. 17/746,848,18/095,463, and 18/197,092.

BACKGROUND

The present invention proposes methods and apparatus for improving theefficiency of heat transfer between two fluid streams, particularly assaid methods and apparatus relate to improving the technology disclosedin U.S. Pat. Nos. 3,225,538, 3,067,594, 3,871,179, Pending U.S. patentSer. Nos. 18/095,463, and 18/197,092, wherein techniques are detailedfor creating, among other useful methods and apparatus, a uniquethermochemical cycle, termed the Bland/Ewing Cycle (B/E Cycle) after theco-inventors behind U.S. Pat. No. 3,225,538, involving “molecularexpansion” and “molecular compression”. The advantage of the B/E Cycleis best exemplified in FIG. 3 and FIG. 4 of U.S. Pat. No. 3,225,538,where PN and T/S charts indicate the potential for increased “powerdensity”. This power density is a result of the reduced compression workin (W-in) following exothermic conversion to fewer moles of gas relativeto the increased expansion work out (W-out) following endothermicconversion to increased moles of gas.

This invention particularly relates to improvements to methods andapparatus that permit the efficient employment of endothermic chemicalreactions and reversible chemical reactions of theendothermic-exothermic type for transfer of heat and/or production ofmechanical energy.

The underlying foundational invention takes the form of a unique heattransfer system or STREP in which a counter-flow regenerator canuniversally replace a counter-flow recuperator to good effect,particularly where a STREP can increase the efficiency when anendothermic chemical reaction and/or an exothermic chemical reaction areutilized. In one aspect, the heat transfer method or system of thisfoundational invention may be adapted to heat a space or a substance orit may be embodied as a refrigeration system. In another aspect, theheat transfer method may take the form of a method or system for theproduction of mechanical work. In the application to thermochemicalprocesses, endothermic and exothermic methods or systems may becyclical, wherein a chemical reactant endothermically reacts to form aproduct or products and the product or products are then reacted tore-form the initial chemical reactant. Also it is contemplated that areactant which will undergo an endothermic chemical reaction may beemployed to do mechanical W-in to a method which does not involveconverting the products back to the initial reactant. Also it iscontemplated that the reformation to the initial chemical substance,since it evolves the total amount of thermal energy absorbedendothermically, may itself drive heat engine processes that produceW-out, said exothermically-produced W-out then being summable with theW-out produced endothermically to equal a total or net W-out for acomplete Bland/Ewing cycle. Also it is contemplated that saidreformation to the initial reactant can be designed to primarily producethermal energy rather than W-out. Also it is contemplated that theendothermic process may be designed to primarily produce cooling bysubstantially lowering the temperature of the product of endothermicdissociation prior to expansion. Also is contemplated that theBland/Ewing Cycle, when viewed as composed of two half-cycles, can beseen as an efficient means of transporting hydrogen in liquid form atambient pressure and temperature via a process termed a Benzene Batteryas described in U.S. patent application Ser. No. 18/197,092.

Stated broadly, this foundational invention utilizes the improvedcharacteristics of what might be broadly termed a “valved regenerator”over the characteristics of a standard counter-flow recuperatorvis-a-vis increasing the efficacy of heat transfer between two fluidstreams, particularly as concerns the two fluid streams associated witha Bland/Ewing Thermochemical Cycle.

This background section is provided only for purposes of introducingcertain background material relating to the present disclosure and,thus, is not an admission of prior art.

SUMMARY

In several embodiments of the STREP method proposed herein, the efficacyof replacing a counter-flow recuperator with a valved regeneratorconcerns the ability to efficiently change the temperature of twocounter-flowing streams with markedly reduced internal volumes. This iswell known to benefit what are termed “stirling engines”, which flow afixed quantity of fluid back and forth between two constantly changingvolumes through an intermediate thermal sponge or regenerator. The STREPconcept, however, perceives the fluid flow as being composed of twodifferent streams, which only incidentally may be composed of a fixedamount of common fluid within some device. This makes it particularlyuseful when exchanging thermal energy between an endothermic fluidreactant and an exothermic fluid product.

The STREP also is differentiated from a stirling engine regenerator inbeing capable of flowing fluid through a regenerator with setparameters, including isobaric (constant pressure), isochoric (constantvolume), isothermal (constant temperature), and all the possibilities inbetween. It has even been found to be capable of flowing one stream withone set of parameters, as for example isobaric, and the second streamwith another set of parameters, as for example isochoric.

This Summary section introduces some features of non-limiting andnon-exhaustive examples of the present disclosure, and is not intendedto limit the scope of the claims.

BRIEF DESCRIPTION OF THE DRAWINGS

The drawings, together with the specification, illustrate non-limitingand non-exhaustive example embodiments of the present disclosure.

FIGS. 1, 2, 3, and 4 each illustrate a cross-sectional view of asynchronized thermal regenerator exchange pump (STREP) during differentstages of a synchronized heat exchanger process according to someexamples.

FIGS. 5A and 5B schematically illustrate a STREP system according tosome examples.

FIGS. 6A and 6B schematically illustrate a STREP system according tosome examples.

FIG. 7 is a photograph of a Closed Cycle Valved Cell (CCVC) prototypeengine.

FIG. 8 illustrates a cutaway view of a solid model upon which the CCVCprototype engine was based.

FIG. 9 illustrates a solid model of the multi-spiraled helical ribbingwhich constituted the flow channels for the various CCVC heatexchangers.

FIG. 10 illustrates how the CCVC prototype heat exchangers weregenerally constructed of multi-spiraled helical ribbing, whichconstituted the flow channels for the various CCVC heat exchangers.

FIG. 11 illustrates an isometric and cross-sectional view into anexample Synchronizing Displacer (SD) CCVC (SD-CCVC) design.

FIG. 12 illustrates a front view of the SD-CCVC design of FIG. 11 at TopDead Center (TDC).

FIG. 13 illustrates a front cross-sectional view of the SD-CCVC designof FIG. 11 at a 90 degree crank rotation following TDC.

FIG. 14 illustrates a front cross-sectional view of the SD-CCVC designof FIG. 11 at bottom dead center (BDC).

FIG. 15 illustrates a front cross-sectional view of the SD-CCVC designof FIG. 11 at a 90 degree crank rotation following BDC.

FIG. 16 illustrates a front cross-sectional view of a M-SD design at a90 degree crank rotation following TDC according to an example.

FIG. 17 illustrates a front cross-sectional view of the M-SD design at a90 degree crank rotation following BDC according to an example.

FIG. 18 illustrates a pressure/volume/temperature/energy/entropy chartbased on FIG. 70 , “Marks Mechanical Engineers' Handbook”, 1st edition,9-148, “Internal-combustion engines.”

FIG. 19 illustrates a separated-out tracing of working fluid statesutilizing the chart illustrated in FIG. 18 .

FIG. 20 illustrates estimated pressure and volume information from FIG.19 in graphed curves.

FIG. 21 illustrates chart 14 of the July 2005, EISG report.

FIG. 22 illustrates a graph based on FIG. 1 of U.S. Pat. No. 3,225,538.

FIG. 23 illustrates a simple system for storing and delivering C6H12.

FIGS. 24 a, 24 b, and 24 c illustrate a combined liquid C6H12 and liquidC6H6 storage tank system.

FIG. 25 illustrates a simple heat generating system, as shown in FIG. 8of U.S. patent application Ser. No. 18/095,463.

FIG. 27 illustrates certain traced, separated-out lines using FIG. 18 .

FIGS. 29 a and 29 b each illustrate a cutaway view of a CAD solid modelof an approach to converting an existing CCVC prototype engine accordingto some examples.

FIGS. 30 a, 30 b, and 30 c each illustrate a cutaway view of a CAD solidmodel of an approach to converting an existing CCVC prototype engineaccording to some examples.

FIG. 32 illustrates a mixed isobaric/isochoric STREP according to someexamples.

FIG. 33 illustrates an example cycle according to some examples.

FIG. 34 illustrates a low pressure cylinder and a high pressure cylinderof a STREP according to some examples.

FIGS. 35, 36, 37, and 38 illustrate working fluid pathways for workproducing cycles and a refrigeration cycle using certain traced,separated-out lines using FIG. 18 .

FIG. 39 is copied from FIG. 25 and labels a simple exothermic heatgeneration system schematic with labels from FIG. 27 .

FIG. 40 illustrates FIG. 2 from U.S. patent application Ser. No.18/197,092.

FIG. 41 illustrates FIG. 40 reconfigured as a refrigeration cycleaccording to some examples.

FIG. 42 illustrates a mixed isobaric/isochoric STREP according to someexamples.

DETAILED DESCRIPTION

Adding an SD mechanism to a regenerator is herein proposed as a meansfor greatly increasing the overall efficiency of any heat exchangeprocess. A regenerator is a “thermal sponge” that absorbs thermal energyfrom a working fluid when it flows in one direction and releases thatthermal energy back to the working fluid in the reverse direction. Abasic STREP would essentially be composed of a receiver cylinder andpiston means, an SD cylinder and piston means, valving and connectingmanifold means, and a regenerator. The STREP heat exchanger processwould function as follows and as illustrated in solid modeled andcross-sectioned FIG. 1 through FIG. 4 below (Note: dashed lines andarrows illustrate fluid flow in FIG. 1 through FIG. 4 ):

-   -   (1) See FIGS. 1 and 2 . Through an intake valve (In1),        intermittently pass a stream of working fluid at constant        pressure (P1) and at temperature (T1) through a regenerator (A),        through a receiver mechanism intake valve (In2) and into a        receiver mechanism such as a piston-and-cylinder arrangement        (B), thus changing the temperature of the working fluid in the        receiver mechanism to a different temperature (T2) via        conduction of thermal energy either into or out of the material        of the regenerator. Simultaneously, through an intake valve        (In3), intermittently pass a second working fluid stream from        some external system at constant pressure (P2) and temperature        (T3) into a second “synchronizer mechanism”, such as a SD        piston-and-cylinder arrangement (C).    -   (2) See FIGS. 3 and 4 . At constant pressure, intermittently        pass the first working fluid stream at P1 and T2 out of the        receiver mechanism piston-and-cylinder arrangement, through a        receiver mechanism exhaust valve (Ex1), and into some external        system. Simultaneously, at constant pressure, intermittently        pass the second working fluid stream at P2 and T3 through an        exhaust valve (Ex2), out of the synchronizer mechanism SD        piston-and-cylinder, through a regenerator intake valve (In4),        through the regenerator, through an exhaust valve (Ex3) on the        opposite side of the regenerator near intake valve In1, and into        some external system, thus changing the temperature of the        working fluid in the receiver mechanism to a temperature that        approaches T1 via conduction of thermal energy either into or        out of the material of the regenerator.

Since a constant pressure is maintained within the receiver cylinderduring intake and exhaust, receiver cylinder W-in cancels W-out,reducing any W-in to that required to overcome any pumping losses. For asimilar reason, W-in and W-out for the SD cylinder also cancels outexcept for pumping losses.

Arrows with solid lines within FIGS. 5 a through 6 b illustrate fluidflow through connecting manifolding/ducting. Dashed lines illustratedisconnected manifolding/ducting and no fluid flow. For continuous flow,two or more regenerators operating intermittently and in syncopation canbe used, one or more filling while the others are depleting. Theschematics in FIGS. 5 a and 5 b show such a system, in this case showingthe use of a single fan, two regenerators and a series of valves as analternative to positive displacement mechanisms. The use of a fan orturbine is particularly useful where the regenerator is relatively largecompared to the flow rate. Thus, for example, the exhaust from a fireheating a residence can be blown through a large regenerator from theregenerator “hot” side to the “cold” side, absorbing thermal energy,then be exhausted to the outside atmosphere. When the regenerator isthermally “full”, then clean air, for example from within the residence,can be blown through the regenerator in the opposite direction,exhausting back into the residence and heating the residence with thethermal energy deposited in the regenerator earlier.

The process shown in FIGS. 5 a and 5 b can be shown to operate in thefollowing manner (see Table 1 below):

Table 1

-   -   A—Fan    -   B—Inlet air stream two way splitter    -   C—Heat source    -   D—Two way valve #1    -   E—Two way valve #2    -   F—Regenerator #1    -   G—Two way valve #3    -   H—Heat source exhaust    -   I—Two way valve #4    -   J—Two way valve #5    -   K—Regenerator #2    -   L—Two way valve #6    -   M—Heated clean air exhaust

In FIG. 5 a:

-   -   A. A fan receives clean air, as from within a residence.    -   B. The clean air stream produced by the fan enters a two way        stream splitter.    -   C. One of the two air streams feeds into a heat source, in this        case an enclosed fireplace.    -   D. The hot exhaust from the fireplace passes through two way        valve #1, where it is directed to two way valve #2.    -   E. Two way valve #2 directs the hot exhaust to regenerator #1.    -   F. The hot air passes through regenerator #1, charging the        regenerator with heat, and the hot air being cooled in the        process.    -   G. The cooled air passes through two way valve 3.    -   H. The cooled air is exhausted, in this case outside of the        residence.    -   I. Simultaneously, the second stream of clean air proceeding        from the fan and through the 2 way splitter (see step B, above)        is directed to two way valve #4, which directs the clean air to        two way valve #5.    -   J. Two way valve #5 directs the clean air to previously        thermally charged regenerator #2.    -   K. The clean air passes through regenerator #2, cooling the        regenerator and the cooling air thus being heated.    -   L. The heated clean air passes through two way valve #6.    -   M. Finally, the heated clean air is passed back into the house.

In FIG. 5 b:

When regenerator #2 has been “emptied” of its thermal energy, the twoway valves #1 through #6 are switched to the alternative setting. Nowhot exhaust from the fireplace passes through two way valve #1 (D),through two way valve #6 (L), through regenerator #2 (K), through twoway valve #5 (J), and finally exhausts outside the residence (H).Simultaneously, a stream of clean air proceeding from the fan (A),through the two way splitter (B), through two way valve #4 (I), throughtwo way valve #3 (G), through regenerator #1 (F), and finally exhaustsinside the residence through two way valve #2 (E).

As an alternative to the valving arrangement shown in FIG. 5 , the tworegenerators shown in FIG. 5 can be made to intermittently exchangeplaces, as by a rotation around their center, such that the position ofregenerator #1 (F) is on the top and the position of regenerator #2 (K)is on the bottom, allowing the stream of air cooling the regenerator toalways be directed to the top, and the stream of air heating theregenerator to always be directed to the bottom, thus simplifying thevalving. Essentially, FIG. 5 b would look exactly like FIG. 5 a and thedotted lines representing an intermittent change in stream directionwould be eliminated and replaced by an intermittent movement ofregenerator K to the position of regenerator F and regenerator F to theposition of regenerator K.

It is also possible, where input heat could be turned on or offintermittently, to use a single non-moving regenerator (see theschematics in FIGS. 6 a and 6 b ), to intermittently heat and cool theregenerator. In that case, the exhaust from fan A would be replaced withtwo way valve I that either (1) feeds fire C and exhausts throughtwo-way valves D and E, through regenerator F, through two way valve G,exhausting outside with cooled combusted air H, or (2) heats the houseby passing from fan A through two way valve I, through two-way valve G,back through regenerator F, through two-way valve E and exhausting intothe residence with clean heated air M.

A second two way valve (not shown) may be used to select whether airflows to the fan from inside the house or from outside the house. Sinceair from inside the house would generally be warmer than air outside thehouse, drawing air for heating the house from the house helps maintainheated air within the house. Air within the house thus essentiallyrecirculates through the regenerator. Note that, if air from the housewere used to feed the fire, which is then exhausted outside the house,“makeup” cold air from outside would need to be drawn into the housefrom somewhere to adjust for the air being removed to feed the fire.

SD Heat Engines

SD heat engines are proposed herein, including variants. The SD conceptis predicated on the breakthrough concept of adding a “synchronizedthermal regenerator exchange pump” or STREP to the original externallyheated Closed Cycle Valved Cell (CCVC) heat engine concept (as conceivedand constructed under a California Energy Commission (CEC) EnergyInnovation Small Grant (EISG) early in the 21st century. Thismodification would create an “SD-CCVC” engine. An SD permits the CCVCprocess to utilize a “regenerator”, significantly improving powerdensity and overall real-world efficiency. A regenerator is a “thermalsponge” that absorbs and releases thermal energy from and to a workingfluid. A well-known engine that uses a regenerator is a stirling engine,which is roughly based on the Stirling Cycle. A regenerator can absorbheat available in a stirling engine's working fluid following workingfluid expansion and release a substantial amount of that thermal energyback to the working fluid prior to the addition of a charge of “new”thermal energy from an outside heat source, thus reducing the amount of“new” thermal energy required to operate the engine. A regenerator canalso remove thermal energy with a charge of “cold” from a cold source.This allows stirling engines to essentially be run in reverse, creatingrefrigerated working fluid when the fluid is expanded to below ambienttemperature. Such a refrigerating process requires a work source tocompress the working fluid and overcome pumping losses. Note that anSD-CCVC engine can also operate as either a heat engine or as arefrigerating engine, although valving would have to be extensivelymodified.

As will be shown, there are several possible variants other than anSD-CCVC heat engine. These include:

An Open Cycle or OCVC or SD-OCVC heat engine that is externally heatedand uses compressed air as the working fluid.

An OCVC or SD-OCVC heat engine that is heated by internal combustion(i.c.) and uses compressed air as the working fluid.

A Mixed heat source or M-OCVC, or M-SD-OCVC heat engine utilizingmultiple sources of heat, such as solar heat (medium temperature), heatfrom the external combustion of a fuel (high temperature), and/ori.c.-derived heat (very high temperature).

An OCVC, M-OCVC, SD-OCVC or M-SD-OCVC heat engine with internal heat in(H-in) produced by injection and i.c. of a fuel and an oxidant into acompressed gas or vapor prior to and/or during the expansion process,where the main body of working fluid into which the fuel and oxidant areinjected and combusted would be air, where the combusted products plusair are essentially completely removed at the end of each cycle, and anew charge of air is taken in.

A CCVC, M-CCVC, SD-CCVC or M-SD-CCVC heat engine with internal H-inproduced by injection and i.c. of a fuel and an oxidant into acompressed gas or vapor prior to and/or during the expansion process,where the combusted products are essentially completely removed at theend of each cycle, such as by liquefaction of H2O, where the main bodyof working fluid into which the fuel and oxidant are injected andcombusted would be non-reactive, such as He, and where the main body ofworking fluid is continually recirculated.

A “Benzene Battery” (BB) or BB-CCVC, BB-SD-CCVC, or BB-M-SD-CCVC heatengine, where the fuel is H2 delivered by a cyclical hydrocarbon such asC6H6 (benzene) and thus the cyclical hydrocarbon is completely recycled.(See “The BB Closed Loop Process” below.)

A BB-CCVC, BB-SD-CCVC, BB-M-SD-CCVC, where the H2 is from a BB and theoxidant is compressed O2 gas, O2 liquid, or O2 released from a chemicalcarrier such as H2O2, and both the H2O and the cyclical hydrocarbon suchas C6H6 are completely recycled. (See “The BB Closed Loop Process”below.) Such H2+O2-burning engines may be characterized as part of aclosed-cycle energy capture and conversion system, the heat engine beingsupplied H2 by the endothermic dissociation of a cyclical hydrocarbonsuch as cyclohexane (C6H12) into a “carrier hydrocarbon” such as C6H6,and the heat engine being supplied O2 in either compressed gas form,liquid form, or chemical form such as H2O2, where said H2 and O2 arecontinually recycled in the form of easily stored and shipped exhaustedH2O and C6H6, which are potentially continually reusable. For example,such a process put in place on the lunar surface would ideally onlyrequire the original chemical constituents, the mechanisms themselves, asource of high temperature source energy such as concentrated solarenergy, a means for removing waste heat such as a radiator or cooler,and various storage and shipping means. Note that all the oxidizer andfuel chemical constituents can be stored indefinitely, thus acting as akind of “battery” for releasing thermal energy over the two week longlunar night.

The Existing CCVC Design.

As mentioned earlier, FIG. 7 is a photograph of the existing CCVCprototype. FIG. 8 is a cutaway of a solid model upon which the CCVC wasbased. FIG. 9 is a solid model of the multi-spiraled helical ribbingwhich constituted the flow channels for the various CCVC heatexchangers. FIG. 10 is a photograph showing how the CCVC prototype heatexchangers were generally constructed, in this instance showing fivemulti-spiraled ribbed channels. If the multi-spiraled ribbed channelillustrated in FIG. 10 were inserted within a second larger diametermulti-ribbed channel, that would create two counter-flowingmulti-channel fluid streams passing in close physical proximity to oneanother, as is shown in FIG. 9 , thus allowing two counter-flowing fluidstreams to exchange heat with one another.

The existing CCVC prototype, as shown in FIG. 8 , uses a singlecontiguous hollow piston which creates four distinct chambers within twoequal-sized cylinders connected by a smaller diameter third cylinder;the expansion chamber, the compression chamber, the lower displacerchamber, and the upper displacer chamber. These were created by placingtwo piston discs or heads of 2.5″ diameter on either end of a 2″diameter connecting tube. Running teflon seals were added to theperimeters of the two heads, and a third stationary teflon seal was incontact with the 2″ diameter connecting tube. A small 0.375″ diameterguided drive rod, also put in contact with a teflon seal at the base ofthe compression cylinder, connected to a piston guide captured betweenguide blocks, said piston guide being driven by a standard pistonconnecting rod connecting to a standard crankshaft with a throw of2.75″. At Top Dead Center (TDC), the (bottom) compressor space wouldthus have a volume of 13.39 cu in. At Bottom Dead Center (BDC), the(top) expander space would have a volume of 13.5 cu in. At TDC, thespace between the 2.5″ diameter expander cylinder wall and the 2″diameter connecting tube created a space in the upper displacer cylinderwith a volume of 8.64 cu in. At BDC, the space between the 2.5″ diametercompressor cylinder wall and the 2″ diameter connecting tube likewisecreated a space in the lower displacer cylinder with a volume of 8.64 cuin.

The existing CCVC prototype is designed to use teflon+stainless steelspring seals for its piston rings and tube and guided drive rod seals,permitting essentially non-lubricated, low friction movement of the CCVCpiston. The internal engine volumes are pre-pressurized to some desiredpressure.

In action, beginning at ˜TDC, a cold working fluid such as helium atsome pressure is drawn through a poppet intake check valve into thelower displacer cylinder. At ˜BDC, the working fluid is exhausted fromthe lower displacer cylinder through a poppet exhaust check valve. Theworking fluid passes into the recuperator counterflow heat exchanger(shown in outline as dotted lines in FIG. 8 ), from lower to upper andfrom cold to hot. The working fluid cyclically passes through an innermulti-spiraled set of helical ribbing, where it is preheated withworking fluid cyclically exhausting from the engine expander through acompletely separate outer multi-spiraled set of helical ribbing. Thefluid exits the hot end of the heat exchanger and passes into the top ofthe heater heat exchanger.

In the existing CCVC prototype, the heater was electrically heated by aninternal cartridge heater and an external coil heater. The heater heatexchanger, like the recuperator, has an inner and outer multi-spiraledset of helical ribbing. Note that the heater was originally constructedwith seals such that the inner multi-spiraled set of helical ribbingreceived working fluid from the recuperator on its way to the upperdisplacer, and the outer multi-spiraled set of helical ribbing receivedworking fluid from the upper displacer on its way to the expander. Theelectric cartridge heater was put in physical contact with the innerribbing and the electric coil heater was put in physical contact withthe outer ribbing. However, the design was changed, in an attempt toreduce pumping losses, to run working fluid from the recuperator throughboth ribbed spirals and into the upper displacer during the upstroke,and through both ribbed spirals and into the expander during thedownstroke. Note that doing so had zero impact on the total volume“seen” within the heater by working fluid.

The displacement between the lower displacer and the upper displacer istherefore a constant volume waste heat addition process that completesat ˜TDC and thus raises the pressure of the captured working fluid.

Slightly before TDC, a special poppet-type “transfer valve” (expanderintake valve) is opened that connects the working fluid in the upperdisplacer, recuperator, and heater to the expansion chamber. The valveis actively biased to automatically open, such that, if pressures onboth sides of the poppet head are equal, the valve will pop open. Toequalize pressure on both sides of the transfer valve poppet head, thepoppet-type expander exhaust valve, which is biased towards closed andmechanically driven open by a rocker arm connected to a push rodconnected to a cam on the crankshaft, closes slightly before TDC. Deadspace at TDC is minimized, which allows any remnant working fluidcaptured in the expander at the close of the expander exhaust valve (asit approaches TDC) to be re-pressurized to at or above the pressure ofthe hot, pressurized working fluid in the upper displacer and theheater. Consequently, the transfer valve wants to pop open. A tinyprojection at the bottom of the transfer valve poppet head is designedto physically contact the top of the piston just prior to TDC, thusensuring that the transfer valve will in fact begin to open.

As the expander travels from TDC to BDC, the upper displacer then isable to exhaust the working fluid back through the heater, through thetransfer valve, and into the expander. Since the displacer is low volumeand the expander is high volume, an expansion process thus occurs. Notethat the expanding working fluid includes the volume captured in therecuperator, the upper displacer, the heater, and the various manifoldsand plenums connecting these elements.

As the piston approaches BDC, the exhaust valve begins to open into theexpander cylinder, where the working fluid pressure has been reduced byvolumetric expansion. At the same time, an arm on the exhaust valvepushes the transfer valve towards closed and holds it in place there.Note that, as pressure builds within the upper displacer cylinder duringthe ensuing displacement “charging” stroke, the transfer valve willeventually hold itself closed by pressure differential. Thus, when theexhaust valve begins to close as it nears TDC, it lifts the arm off ofthe transfer valve, leaving it prepared to automatically pop open againwhen pressure access the transfer valve head equalizes in the mannerdescribed above.

From BDC, the expander piston now increases pressure, until it reachessufficient pressure to drive the expanded working fluid past a lightlybiased check valve. (Note: It is surmised that one of the major reasonsthe prototype was unable to produce net W-out was due to a very earlyclosure of the transfer valve, which caused super-expansion andrecompression in the expander. The super-expansion is the reason for theaddition of the exhaust check valve.) This pressuredifferential-actuated check valve is constructed integral to themechanically operated exhaust valve, essentially sliding back and forthalong the valve stem. With the opening of the exhaust check valve, theexhausting working fluid is allowed to enter the outer multi-spiraledhelical ribbing of the recuperator, thus passing otherwise-waste heat tothe inner multi-spiraled helical ribbing, as described above.

The exhausting fluid from the expander then exits the recuperator andenters the multi-spiraled helical ribbing of the cooler, which furtherdrops the working fluid temperature. Finally, the exhausting fluidenters the compressor cylinder near the base of the engine that sits ontop of the drive unit. Since the volume of the compressor cylinder atfull extension closely approximates the volume of the expander at fullextension, the process of moving fluid out of the expander and into thecompressor essentially occurs at constant volume, albeit a tiny amountof work is required to overcome the small volumetric difference. (Note:This volumetric difference can be avoided by passing a rod of exactly0.375″ diameter out of the top of the expander. However, it wouldrequire a high temperature seal contacting that rod plus some method oflubrication.)

The process of exhausting from the expander, through the two heatexchangers, and into the compressor can thus be seen as a kind ofconstant volume displacement process that removes heat from theexhausting working fluid. Since this heat removal occurs at constantvolume, the pressure of the working fluid during exhaust is thuscontinually reduced until TDC is reached.

Following TDC, the direction of the compressor piston is reversed, andthe volume composed of the interior of the recuperator, the cooler, andthe compressor cylinder begins to climb in pressure from the resultingmechanical compression. (Note: In the original design, the exhaust fromthe compressor was directed to the displacer intake check valve.) Alsoat TDC, the volume in the lower displacer cylinder assembly will beginto expand, automatically closing the lower displacer poppet exhaustvalve, which is lightly biased towards closed to reduce pumping lossthrough the valve. That will permit a fresh charge of pressurized andcooled gas to be taken into the lower displacer cylinder, thus returningthe engine to its initial state at TDC and completing a full cycle.

Testing of the original CVCC prototype verified that the expectedpressure differentials were in fact occurring as predicted. However, nonet W-out was ever observed in the existing CCVC prototype. In largedegree, that was determined to be the result of four factors:

-   -   (1) Pre-pressurization of the existing CCVC prototype was too        low to generate sufficient power density to overcome the        engine's friction and pumping losses.    -   (2) The large internal volumes of the heater+recuperator and the        cooler+recuperator greatly reduced the potential pressure        differentials.    -   (3) Too-early closure of the expander transfer valve.    -   (4) the likelihood of a too limited temperature spread in        comparison to the fixed displacer/expander volume ratio.

Regarding (1), since the prototype required W-in to rotate, a higherpressure was difficult to achieve while the prototype was being rotatedwith no net W-out during the process. That problem would be solved ifnet W-out could be increased.

Regarding (2), the SD-CCVC design, in replacing the large internalvolume recuperator with a much smaller internal volume regenerator,would solve that problem.

Regarding (3), it is likely that, since the pass through volume from theupper displacer into the expander was at a maximum halfway through thestroke, a “suction” was developed that helped to unseat the transfervalve, due to the maximum speed of the dropping expander piston beingachieved at exactly that halfway point. Therefore, converting thetransfer valve to a fully physically-actuated valve rather than apartially pressure differential-actuated valve should solve thatproblem.

Regarding (4), increasing the relative volume of the displacer pistonwould create a better match due to a decreased displacer-to-expanderexpansion ratio.

The SD-CCVC design being proposed herein is expected to address all but#3 of these issues.

One Possible SD-CCVC Heat Engine Design.

FIGS. 11 through 15 indicate one possible SD-CCVC heat engine design, inthis case utilizing most of the elements of the existing CCVC prototype,albeit rearranged, partially modified, and/or duplicated. The proposedSD-CCVC design, being closed cycle, requires that all source heat beadded by some form of external heat exchanger (external heater).

FIG. 11 is an isometric and cross-sectioned view into the proposedSD-CCVC design, and labels various elements of the proposed SD-CCVCdesign. FIG. 12 is a front view of FIG. 11 at TDC, and additionallylabels various elements. FIG. 13 is a front cross-sectioned view at a 90degree crank rotation following TDC. FIG. 14 is a front cross-sectionedview view at BDC. FIG. 15 is a front cross-sectioned view at a 90 degreecrank rotation following BDC (a note in FIG. 15 illustrates the locationof a potential H2O removal site, whose usefulness will be madeapparent).

Table 2 below describes and defines the parts of the proposed SD-CCVCdesign as shown in FIG. 12 . Table 2 is also referred to under theheading below entitled “Cycle analysis”.

Table 2

-   -   A. Lower displacer-and-2nd stage compressor cylinder    -   B. Regenerator (STREP)    -   C. Upper displacer cylinder    -   D. Displacer piston connecting tube    -   E. Lower displacer actuated exhaust valve    -   F. Lower displacer-and-2nd stage compressor piston    -   G. Lower displacer intake check valve    -   H. Expander cylinder    -   I. Expander piston    -   J. Expander exhaust valve    -   K. Expander intake transfer valve    -   L. External heater    -   M. Upper displacer piston inlet and regenerator exhaust check        valve    -   N. Upper displacer piston    -   O. SD cylinder    -   P. Guided piston rods and guides (3 ea) (guide blocks not shown)    -   Q. SD cylinder actuated inlet valve    -   R. SD cylinder exhaust check valve    -   S. SD piston    -   T. 1st stage compressor-to-SD connecting rod    -   U. 1st stage compressor intake transfer valve    -   V. 1st stage compressor cylinder    -   W. 1st stage compressor piston    -   X. 1st stage compressor cylinder actuated exhaust valve    -   Y. 1st stage compressor cylinder exhaust check valve    -   Z. 1st stage compressor exhaust cooler    -   AA. 2nd stage compressor piston connecting tube    -   AB. 2nd stage compressor intake check valve    -   AC. 2nd stage compressor exhaust check valve    -   AD. 2nd stage compressor exhaust cooler    -   AE. SD inlet check valve

The existing CCVC prototype has (from bottom to top) the compressor, thelower and upper displacer, and the expander in a singlevertically-combined assembly operated by a single crank throw. Theproposed SD-CCVC design, has three assemblies (see FIG. 11 ) made up ofan expander assembly (on the left), a vertically-combined displacer and2nd stage compressor assembly (in the middle), and a vertically-combinedSD and 1st stage compressor assembly (on the right). It is obvious thatother ways to combine the various elements are possible. Finally, notethat all three crank throws are in line and perfectly syncopated, thussimultaneously hitting both their respective TDCs and BDCs. Note that,though not shown in this configuration, it is possible for a singlecrank throw to drive all three assemblies. That may be useful in lettingthe three separate assemblies be “wrapped around” a central drive rod ina future iteration, thus minimizing manifold spacing.

The SD-CCVC heat engine design shown in FIGS. 11 through 15 can be saidto vary in large part from the basic design of the existing CCVCprototype by the addition of an SD cylinder (O) and piston (S). In thisiteration, the added SD cylinder and piston match the bore of theexpander cylinder (H) and the stroke of the expander piston (I), butoperate 180 degrees out of phase to the expander piston. Therefore,since the working fluid exhausted from the expander cylinder istransferred to the SD cylinder at essentially constant pressure,temperature and volume via a manifold (not shown), only pumping lossesand thermal losses need to be taken into consideration as affecting thestate of the working fluid before and after the transfer.

A second way the proposed SD-CCVC design varies from the existing CCVCprototype concerns the proposed means of recapturing otherwise-wasteheat. The existing CCVC prototype relied on classic counterflow externalheat exchangers at various points in the cycle. In contrast to theexisting CCVC design, a STREP, which is a kind of internal thermalregenerator (B), is used in the proposed SD-CCVC design. The use of aSTREP is effectively made possible by adding the SD to the existing CCVCengine design.

A STREP's use of internal thermal regeneration makes it more compactthan counterflow thermal exchange/recuperation, and thus more practicalfor a cycle that relies on constant volume displacement processes, suchas the existing CCVC engine. Thermal regeneration is also generally moreefficient for heat transfer than thermal recuperation, since itincreases the ability to more completely transfer the total heatdifferential between the counter-flowing streams of working fluid.

In addition to the SD and the STREP, a two stage inter-cooled compressorsystem has also been added to the proposed SD-CCVC design. Increasingthe number of inter-cooled compression stages is well known to assist inapproaching an isothermal compression, which will aid overall thermalefficiency.

The 1st stage compressor piston (W) is designed to match the diameterand stroke of the SD piston. Thus, having passed through the STREP, the1st stage compressor cylinder will receive the working fluid exhaustedfrom the SD cylinder at constant volume. To accomplish this, the 1ststage compressor piston has a small diameter 1st stage compressor-to-SDconnecting rod (T) on the opposite side of the piston head from thedrive unit. The connecting rod passes through a stationary teflon sealheld in the 1st stage compressor cylinder head and attaches to the SDpiston head, causing (1) the SD piston to exhaust a charge of workingfluid through the STREP and into the 1st stage compressor cylinder whenthe 1st stage compressor piston takes in working fluid, and (2) the SDpiston to take in a fresh charge of working fluid from the expansioncylinder's exhaust manifold when the 1st stage compressor space exhaustsits latest charge of working fluid. Thus, this working fluid exchangeprocess from the SD cylinder via the STREP to the 1st stage compressorcylinder occurs at essentially constant volume.

The proposed SD-CCVC design's lower (A) and upper (C) displacercylinders operate exactly like the existing CVCC lower and upperdisplacer cylinders. Note that the upper displacer cylinder is exactlythe same diameter as the lower displacer-and-2nd stage compressorcylinder, and in this instance is also the same as the expandercylinder, the SD cylinder, and the 1st stage compressor cylinder. Thatis, a large diameter connecting tube (D) forms the inner surface of boththe lower and upper displacer cylinders, and a piston head on either endof that tube (N, F) carry teflon piston seals. As in the existing CVCCprototype, a stationary teflon sealing ring (not labeled) is used on thelarge diameter connecting rod (D), separating the working fluid in thelower cylinder from the working fluid in the upper cylinder. Note: Thestationary displacer cylinder sealing ring is placed appreciably closerto the cooler lower displacer to allow a greater temperature to betolerated in the upper displacer cylinder.

As noted above, the displacer and 2nd stage compressor assembly isphysically separated from the expander cylinder, shown on the left sidein FIGS. 11 through 15 . That separation allows both the expandercylinder and the upper displacer cylinder to utilize a piston with a“standoff extension”, which in turn keeps the teflon piston seal fromrunning on the (hotter) expander cylinder walls by physically moving itfarther away. In addition, separating the two assemblies permits theexpander inlet and upper displacer outlet ports (not labelled) to belocated much closer to one another, reducing manifold space. In likemanner, the SD cylinder, shown on the right in FIGS. 11 through 15 ,will be receiving hot working fluid at constant volume from theexpander, and thus also requires a piston standoff extension that keepsthe piston seal from running on the (hotter) SD cylinder walls. Byphysically removing non-lubricated seals (such as teflon seals) fromrunning directly on the hot expander, upper displacer, and SD cylinderwalls, higher peak temperatures are permitted within the engine, whichincreases potential thermal efficiency. In addition, actively coolingthe portion of the cylinder walls that the non-lubricated seals run onmay permit even higher peak temperatures. It's also possible to activelycool the area surrounding the (non-stationary) piston seals, permittingeven higher peak temperatures.

Note: Because the working fluid passing into the 1st stage compressor(V) will be dramatically cooled and dropped to a lower temperatureduring the displacement expansion of working fluid out of the SDcylinder through the STREP and into the 1st stage compressor, and the2nd stage compressor (A, AA, and F) and lower displacer (A, D, and F)will likewise be receiving dramatically cooled working fluid, the 1stand 2nd stage compressors and the lower displacer will not requirepiston standoffs.

As in the existing CCVC prototype, the lower displacer-and-2nd stagecompressor piston head (F) is double-sided. The 2nd stage compressorcylinder assembly is composed of the lower displacer-and-2nd stagecompressor piston head, the lower displacer-and-2nd stage compressorcylinder (A), and the compressor connecting tube (AA). The 2nd stagecompressor connecting tube connects to the lower displacer-and-2nd stagecompressor piston head on the upper end, passes through a static sealingring (not labeled), and connects to the guided piston rod (P) at thelower end.

Finally, note that “breather holes” (not labelled) are drilled in theplate connecting the lower displacer piston head to the guidedconnecting rod. As a result, the upper displacer piston assembly and thelower displacer-and-2nd stage compressor piston assembly are seen toessentially be composed of a series of connected tubes with varieddiameters. That means the interior of the combined piston assembly caneasily pass a fluid through the interior of the piston via the top ofthe upper displacer piston assembly and the bottom of the 2nd stagecompressor piston assembly “breather holes”. This tube arrangement ispotentially useful for helping internally cool the displacer pistonseals, but also for allowing the elimination of external pressuredifferential across the piston.

Note that, since the area displaced by the upper displacer piston issignificantly larger than the area displaced by the 2nd stage compressorpiston, if one assumes a sealed and constant volume upper and lowercrankcase, then the pressure in the crankcase will elevate during theupstroke of the 2nd stage compressor piston, and reduce during thedownstroke. However, if an inlet check valve were attached to the lowercrankcase and an outlet check valve were attached to the uppercrankcase, then the crankcase fluid would be exhausted at or nearconstant pressure from the upper crankcase via the outlet check valveand would be taken into the lower crankcase through the inlet checkvalve at or near constant pressure in spite of the varying volume. Notethat the fluid thus transferred can be used to help cool the interior ofthe upper displacer piston, again raising the potential peak temperaturethat a non-lubricated piston ring can tolerate.

Seals (not Shown).

The existing CCVC prototype is designed to use teflon+stainless steelspring seals for its piston rings. Teflon seals will function up toabout 555 K (1000 R, 282 deg C., 540 deg F.). However, the SD-CCVCdesign shown in FIGS. 11 through 17 is designed to potentially allowhigher peak temperatures with teflon seals by using standoff extensionsand active cooling.

Estimated Volumes.

Since the proposed SD-CCVC heat engine design based loosely on thedimensions of the existing CVCC prototype engine, volumes can beestimated. Note that, in FIGS. 11 thru 17, many standard elements arenot shown, such as some connecting manifolds, valve springs, etcetera.

Below are volume estimates for the proposed SD-CCVC heat engine designshown in FIGS. 11 through 17 :

-   -   Engine stroke=2.75″ (70 cm).    -   Expansion cylinder (H)=2.5″ (63.5 cm) dia, 4.91 sq in (31.7 cm2)        area*, 13.5 cu in (0.221 L) volume    -   Displacer cylinders (Tot 2) (A, C) volume=2.5″ (63.5 cm) dia,        4.91 sq in (31.7 cm2) area, 13.5 cu in (0.221 L) volume.    -   Displacer piston connecting tube (D)=2″ (5.08 cm) dia, 3.14 sq        in (20.27 cm2) area, 8.64 cu in (0.079 L) volume.    -   Total, displacer cylinder volume minus displacer piston        connecting tube volume (tot 2)=4.86 cu in (0.0796 L).    -   Total, displacer cylinder area minus displacer piston connecting        tube area (tot 2)=1.77 sq in.    -   *Total, expansion cylinder sq in net area=3.14 sq in.    -   SD cylinder diameter and stroke=expansion cylinder diameter and        stroke.    -   1st stage compressor-to-SD piston drive rod (T)=0.225″        (0.572 cm) dia, 0.04 sq in (0.258 cm2) area, 0.11 cu in        (0.0018 L) volume.    -   Total, SD cylinder volume minus SD piston drive rod volume=13.39        cu in (0.219 L) volume.    -   Total, SD cylinder area minus SD piston drive rod area=4.87 sq        in.    -   1st stage compressor cylinder (V)=expansion cylinder volume.    -   1st stage compressor cylinder volume minus SD piston drive rod        volume=13.39 cu in (0.219 L) volume.    -   Total, 1st stage compressor cylinder area minus SD piston drive        rod area=4.87 sq in.    -   2nd stage compressor cylinder=expansion cylinder volume.    -   2nd stage compressor piston connecting tube (AA)=1.58″        (4.016 cm) dia, 1.96 sq in (12.6 cm2) area, 4.86 cu in (0.08 L)        volume    -   Total, 2nd stage compressor cylinder volume minus 2nd stage        compressor connecting tube=8.65 cu in (0.142 L) volume.    -   Total, 2nd stage compressor cylinder area minus 2nd stage        compressor connecting tube area=3.33 sq in.    -   Estimated external heater heat exchanger (L) and manifold        internal volume=1.5 cu in (0.0246 L)    -   Estimated exhaust heat regenerator (STREP) (B) and manifold        internal volume=1.5 cu in (0.0246 L)

Note: The volumes of the expander exhaust manifold, the 1st stagecompressor exhaust manifold, the 1st stage compressor exhaust cooler (Z)and cooler exhaust manifold, the 2nd stage exhaust manifold, the 2ndstage compressor exhaust cooler (AD) and cooler exhaust manifold are notlisted since their volumes are deemed inconsequential to the overallcycle, for the following reasons:

The expander cylinder exhausts working fluid into an insulated manifoldat constant temperature, pressure, and volume which is taken into the SDcylinder at constant temperature, pressure, and volume. Being a constantpressure process, the physical length of the expander cylinder exhaustmanifold is therefore essentially unimportant.

In the cases of 1st and 2nd stage exhaust processes, heat is removedfrom the working fluid as it is being transferred, potentially down tothe temperature of the heat sink. However, the 1st and 2nd stagecompressors are designed to exhaust into their respective manifolds andthrough their respective coolers at approximately constant pressure,since the 2nd stage compressor cylinder assembly and the lower displacercylinder assembly respectively take in working fluid at essentiallyconstant pressure. Therefore, the physical lengths of the 1st and 2ndstage exhaust manifolds are also essentially unimportant.

Finally, it is anticipated that the walls and likely the pistons of the1st and 2nd stage compressors will also be actively cooled. As a result,cooling of the 1st and 2nd stage compressor walls and pistons willassist in helping both compressions approach isothermal. A fulldetermination of the impact of compressor and piston cooling willrequire active testing.

Proposed SD and M-SD Heat Engine Designs.

Proposed SD-CCVC or M-SD-CCVC Designs.

An SD engine would add i.c. source heat only or externally-suppliedsource heat only. An M-SD would add both i.c. source heat and additionalsource heat via an external heater. FIGS. 11 through 17 can be used togenerally illustrate the working mechanisms for an SD or M-SD engine. Inone proposed SD and M-SD design, a means for adding i.c. source heatwithin the expander cylinder (not shown, see “Injector site” in FIG. 11and FIG. 16 and “Alternative injector site” in FIG. 17 ) replaces thesource heat from the proposed external heater shown in FIGS. 11 through17 , and would be located in approximately the same area. Adding a meansfor an i.c. heat source to the SD or M-SD heat engine will allow thecreation of true isothermal, isobaric, or isochoric thermal input, aswell as isobaric, isothermal, and mixed expansion.

In the case of isothermal heat input (H-in), i.c. can maintain aconstant temperature in the expander cylinder throughout expansion, ormay be followed with some amount of adiabatic expansion. In the case ofan isobaric H-in, i.c. can maintain a constant pressure in the expandercylinder throughout expansion, or may be followed with some amount ofadiabatic expansion. In the case of an isochoric H-in, heat would beadded within the expander by near-instantaneous i.c., thus adding heatat essentially constant volume. The expansion that follows can then betailored to anything from isobaric to purely adiabatic by simple timingof a continuing i.c. process, if any.

In the proposed SD and M-SD closed cycle designs, i.c. of H2 and O2 onlywould be arranged, which would produce only H2O. Thus, the exhaustproduct requiring removal each cycle could be composed entirely ofliquid H2O. Assuming the pre-pressurized primary gaseous working fluidto be pure H2, pure O2, any inert gas (such as He), or a mixture ofeither H2 or O2 plus an inert gas, the liquid H2O/H2O2 combustionproduct is easily separated out, with any used H2 or O2 constituent inthe working fluid being continually replenished. Note that such an i.c.engine cycle is defined herein as “closed cycle”, since the product ofcombustion is removed but most of the working fluid is generallyrecycled. One potential site for such a removal is shown in FIG. 15 .

Proposed SD-OCVC or M-SD-OCVC Designs.

In using i.c. heat addition in the apparatus shown in FIGS. 11 through17 , means must be provided for removing the products of combustion.However, with an open cycle design, it isn't possible to do so in themanner undertaken in the closed cycle design, since the products ofcombustion may be primarily gases and not vapors. One possible removalsolution is to exhaust the product of combustion and the working fluid,and take in a new charge of working fluid (for example atmospheric air),as is done with existing Otto Cycle and Diesel Cycle positivedisplacement engines. For use with the engine designs in FIGS. 11through 17 , exhausting of “used air” could be accomplished, forexample, by scavenging out the “used” working fluid (such as air) andreplacing it with “new” working fluid, similarly to the scavengingprocess of existing i.c. two stroke engines. This scavenging would bearranged in the 1st stage compressor cylinder (V), as by arranging theaddition of a second intake valve (for the scavenging air and a secondexhaust valve (for the exhausting “used air”), or possibly a port, thatallowed for scavenging, as is obvious to one skilled in the constructionof 2 stroke i.c. engines. Note that the exhaust could be at or nearatmospheric pressure, or possibly a higher pressure if a supercharger orturbocharger were used.

Alternatively, the 1st stage compressor shown in FIGS. 11 through 17could be used as a displacement receiver only, that simply exhausts the“used air” at or near atmospheric pressure. A separate standard 1ststage air compressor would then be used to inject intercooled air intothe 2nd stage compressor.

Proposed SD and M-SD Design

In the proposed SD and M-SD designs, the means for adding i.c. heatwithin the expander cylinder follows and supplements source H-in fromthe proposed external heater, for example as shown in FIGS. 16 and 17 .One important usefulness of the proposed M-SD process is in potentiallyincreasing the efficiency of converting external heat such as solarenergy into useful work. Recall that the Carnot theorem, or (T1−T2)/T1,where T1 equals the engine's heat source temperature and T2 equals theengine's heat sink temperature, tells us that any process that increasesthe source temperature above a given heat sink temperature will increaseoverall thermal efficiency. Generally, adding i.c.-generated heatfollowing heat addition by an external heater will increase T1, and willtherefore increase thermal efficiency. Thus, even though solar thermalenergy may be added at a lower temperature than T1, the efficiency atwhich that added solar thermal energy is converted into work equals thenet thermal efficiency of the heat engine.

For the proposed M-SD design variant, i.c. would take place followingthe external heater H-in, as shown in FIGS. 16 and 17 . As with the SDvariant, the M-SD variant would be capable of isochoric, isobaric,isothermal, or mixed H-in, and isobaric, isothermal, or mixed expansion.

An Alternative Configuration.

FIGS. 16 and 17 propose an alternative M-SD design to FIGS. 11 through15 . FIG. 16 is comparable to FIG. 13 , and FIG. 17 is comparable toFIG. 15 . The major physical difference is the movement of the externalheater from between the upper displacer cylinder and the expander tobetween the expander exhaust manifold and the SD cylinder.

By positioning the source heater in front of the displacer cylinder, theM-SD process becomes a means for creating what might be termed a kind ofconstant volume thermal supercharger. In function, the source heat beingsupplied by the thermal supercharger is placed “on top” of waste heatcoming out of the expander, thus raising the temperature of the workingfluid transferred into the SD cylinder. On the following SD cylinderexhaust stroke, that higher temperature heat is then temporarily storedin the internal STREP, where it will in turn be transferred into theupper displacer cylinder. In the process illustrated in FIGS. 16 and 17, the thermal supercharger is shown to be an external heater, and i.c.is shown able to add additional heat between the upper displacer and theexpander. Other arrangements for i.c. heat addition are clearlypossible.

As stated, the change in the externally heated M-SD design's expanderexhaust process illustrated in FIGS. 11 through 15 to the internallyheated process illustrated in FIGS. 16 and 17 involves moving from aconstant temperature/pressure/volume displacement process to a changingtemperature/constant volume displacement process. However, the manifoldvolume between the two displacers can play a part in any increase inpressure. If the manifold volume between the expansion cylinder exhaustvalve and the SD cylinder intake valve is relatively large, theexpander-to-SD cylinder process approaches isobaric and the workingfluid temperature within the external heater and in the external heateroutlet increases as the molecular volume increases. If volume isrelatively small, the displacement process approaches isochoric.

In a pure isochoric displacement, during displacement of working fluidfrom the expander cylinder into the SD cylinder, increasing pressurewill raise the temperature of the discrete portions entering the heatexchanger, thus continually reducing the amount of heat per mol eachfollowing discrete portion can absorb as the displacement proceeds.Simultaneously, as the displacement proceeds, for each discrete portionexiting the heat exchanger, the constantly raising pressure of eachprevious discrete portion continually raises the temperature of eachfollowing discrete portion over the temperature originally supplied bythe heat exchanger. As a result, at the end of the displacement process(at TDC), all the working fluid captured within the SD cylinder will beappreciably above the temperature supplied by the heat exchanger. Thatis, a kind of constant volume compression process has occurred. (Note:This is also essentially the case in the proposed SD and M-SD designsfor the working fluid moved from the lower displacer to the upperdisplacer via the STREP. That is, the temperature in the upper displacercould theoretically end up higher than the temperature in the STREP, dueto the internal compression that occurs during displacement.)

Thus, to the degree that the expander-to-SD cylinder process isisochoric, it will raise the pressure in the expander exhaust manifold,external heater, and SD cylinder intake manifold following displacement.That higher pressure needs to be taken into account in the expander'sfollowing exhaust stroke, since the working fluid in the expander at BDCwill want to be at a lower pressure than the working fluid on the otherside of the exhaust valve.

A similar phenomenon occurred with the existing CCVC prototype. In theexisting CCVC prototype, “blow-back” into the expansion cylinder fromthe expander exhaust manifold made it difficult to pump sufficientworking fluid through the engine. As in the CCVC prototype, a checkvalve would need to be added in close proximity to the expander exhaustvalve. In the existing CCVC prototype, a poppet sliding check valve witha light spring bias towards closed was added to prevent blow-back. Thesliding check valve that was developed essentially slid over the valvestem of the expansion cylinder exhaust valve to seat in the cylinderhead, preventing flow backwards into the expander when the expanderpressure dropped below the exhaust manifold pressure. In action,following expansion to BDC, the existing CCVC prototype was allowed tosuper-expand and then recompress its contents until the pressuredifference across the check valve was eliminated, causing the lightlybiased poppet sliding check valve to open, thus permitting the workingfluid in the expander to exhaust.

Such a combined actuated exhaust valve/exhaust check valve design (Y) isillustrated in FIGS. 11 through 17 for use in avoiding blow-back intothe 1st stage compressor, and, while not shown, should be assumed addedto the expander exhaust valve as well for FIGS. 16 and 17 . Note thatother exhaust valve/check valve designs are possible, including whollyactuated designs that don't depend on pressure differential foractuation.

A Possible CCVC Prototype to SD-CCVC Conversion.

Finally, it is possible to “convert” the CCVC prototype to an SD-CCVCprototype by converting the existing compressor to an SD cylinder,adding two external inter-cooled compressors similar to the dualcompressor system shown in FIGS. 11 through 17 , and adding a “valvedregenerator” similar to the valved regenerator system shown in FIGS. 11through 17 . The valved regenerator intermittently and cyclicallyconnects (1) the converted SD cylinder to the 1st stage compressor in amanner that allows the SD cylinder to exhaust at constant volume throughthe regenerator and into the 1st stage compressor similarly to thesystem shown in FIGS. 11 through 17 , and (2) intermittently andcyclically connects the lower displacer to the regenerator in a mannerthat allows the existing lower displacer to exhaust at constant volumethrough the regenerator and into the existing upper displacer similarlyto the system shown in FIGS. 11 through 17 . Note that the existingrecuperator and cooler would in essence be replaced with an insulatedworking fluid transfer tube that connects and intermittently andcyclically creates constant pressure and constant temperature flow fromthe expander exhaust into a valving system added to the existingcompression cylinder similar to the valving system shown connecting theexhaust manifold to the SD cylinder in FIGS. 11 through 17 . Thisworking fluid transfer tube between the expander and the converted SDcylinder would be highly insulated as much as possible to permit the SDcylinder to receive the transferred fluid at constant pressure andtemperature.

Cycle Analysis for the SD-CCVC Design.

FIG. 18 is based on a pressure/volume/temperature/energy/entropy chartfrom FIG. 70 , “Marks Mechanical Engineers' Handbook”, 1st edition,9-148, “Internal-combustion engines”. Basic cycle line drawings such asFIG. 19 will be based on FIG. 18 . For clarity purposes, FIG. 19 andother figures throughout have separated out lines that are traceable onFIG. 18 . Basic cycle line drawings such as FIG. 19 represent thegeneral thermodynamic states of the heat engine working fluid such aspressures, volumes, and temperatures for various heat engine cyclesproposed herein and assist in preparing a first order estimate of thethermodynamic potential of said cycles. Some calculations of estimatedstates of gases will be conducted through analysis of FIG. 18 . However,for additional accuracy, calculations of theoretical states of gaseswill primarily be conducted through use of The Omni Combined Gas LawCalculator (CGL calculator) found at www.omnicalculator.com. Finalefficiency calculations will be based on CGL calculator results. (Note:The existing CCVC prototype was originally designed to allowlow-temperature energy sources to be converted into useful work atdecent thermal efficiencies. Later, the concept was considered for useon the Moon's surface at the lunar poles. A lightweight flat platethermal collector on the lunar surface can easily sustain 600 K (1,080R, 327 C, 620 F) throughout the 2 week long lunar day. Collectingthermal energy in proximity with a Permanently Shadowed Region (PSR) oran artificial PSR, where temperatures in PSR's can approach 100 K (180R, −173 deg C., −280 deg F.), is seen as a means to create a very lowmass solar-powered thermal engine when used in conjunction with a highefficiency low temperature engine such as a CVCC engine.)

The following analysis will assume a working fluid of hydrogen gas (H2),a peak heat engine temperature (T1) of ˜1000 R (555 K, 282 deg C., 540deg F.) and an exhaust temperature (T2) of 400 R (222 K, −51 deg C., −60deg F.). The ideal thermal efficiency of a low temperature engineoperating between a T1 of 555 K and a T2 temperature of 222K is equal tothe well-known Carnot theorem or ((T1−T2)/T1); that is, (555−222)/555,or 60%. Note that real-world engines typically only manage a fraction ofthis theoretical thermal efficiency. In fact, very few if any enginescan operate within a 333 K temperature regime with any meaningfulthermal efficiency at all.

As in FIG. 18 , FIG. 19 temperatures are in degrees Rankine, pressuresare in pounds per square inch, energy and enthalpy are in btu's. Volumesare in cubic feet, but will be corrected for scale, as discussed below.In FIG. 18 , the working fluid is a 1 lb (0.454 kg) mixture of air,atomized fuel, and clearance gas. At 650 R (361 K) and 1 atm (101.3kPa), the mixture occupies about 17 cu ft (481.4 L).

Since an intake volume of 17 cu ft would represent a massive engine,FIG. 18 is clearly representing a much smaller engine undergoingmultiple identical cycles. In the proposed SD-CCVC design, a similarmulti-cycle representation can thus be assumed, based on the mass ofworking fluid per cycle.

In some proposed designs the working fluid would be pre-pressurized,possibly to 1,000 psi (68 atm, 6,894 kPa). Pressurizing the workingfluid, as is done in full scale stirling engines, is a technique forincreasing power density. Note that the impact on thermal efficiency ofpre-pressurizing the working fluid is minimal, but the impact on lossesfrom friction, pumping, etcetera, is highly positive. However, to keepthe information generally contiguous with the boundaries defined by FIG.18 , the proposed SD-CCVC design is assumed to be pressurized to about3.5 atm (˜50 psi, 355 kPa) at an inlet temperature of 222 K (400 R). Itis assumed that doing so will not impact idealized thermal efficiency.

At a temperature of 650 R (351 K, 88 deg C., 190 deg F.), a pressure of1 atm (101 kPa), and a volume of 17 cu ft (481 L) per minute, dry airequals ˜29 g/mol. Assuming 28.5 g/mol to account for water vapor,vaporized fuel (C8H18), and remnant gases, total mol count would equal˜16.9 moles. Using an ideal gas calculator (for example,https://www.meracalculator.com/chemistrv/ideal-gas-law.php), since H2 isnearly n “ideal” gas, H2 volume would equal 17.0 cu ft (481 L) at asimilar pressure, temperature, and mol count (16.6 moles for H2), whichis close to the estimate found for FIG. 19 , and supportive of the useof FIG. 19 as an aid to a first order analysis.

The steps of a typical cycle are described below, with letters in thedescription representing various elements of the proposed design, asdescribed in Table 2 earlier and as described in FIG. 11 and FIG. 12 .The working fluid is assumed to be H2. An analysis of the variouspressure/volume/temperature states theoretically achieved in theseprocesses will indicate the theoretical potential efficiency of thedesign, as traced in FIG. 19 . Note that the cycles shown in FIG. 19 areidealized (no pumping, friction, or thermal leaking losses arepresumed). However, pumping, friction, and thermal losses are expectedto be minimal in a mature design, implying that real world engines ofthis type should approach fairly closely the thermal efficiency of theiridealized cycles.

BDC to TDC—Lower Disolacer to Upper Displacer:

Per the cycle, starting at BDC, the lower displacer cylinder (A) willexhaust, in an isochoric process, cold, pre-pressurized H2 through theregenerator (B) and into the upper displacer cylinder (C). Note that thelower displacer cylinder is composed of the lower displacer cylinderwalls and the displacer piston connecting tube (D), while the upperdisplacer cylinder is composed of the upper displacer cylinder walls andthe displacer piston connecting tube.

To keep the information generally contiguous with the boundaries definedby FIG. 19 , at BDC the lower displacer cylinder H2 working fluid isassumed pressurized to about 3.5 atm (˜50 psi, 355 kPa) at an inlettemperature of 400 R (222 K) (see Point [1] on FIG. 19 ). At BDC, thelower displacer exhaust valve (E) would have been mechanically opened.By TDC, the lower displacer piston (F) would have pumped the H2 workingfluid past the open lower displacer exhaust valve, through theregenerator, past the regenerator exhaust check valve (M) and into theupper displacer cylinder. Since the total volume would have been heldconstant, the H2 working fluid would have increased in both pressure andtemperature. Assuming 100% regeneration of heat, temperature increasedin the displaced mix from 400 R (222 K) in the lower displacer toapproximately the temperature exhausting from the expander, or 720 R(400 K) (see Point [2] on FIG. 19 ).

Per FIG. 19 , at the beginning of source H-in at or near TDC, pressureequals approximately 80 psi (551 kPa). Total volume of the lowerdisplacer, the regenerator, and the connecting manifolds are estimatedto equal about 6.4 cu in (0.1 L). Per the ideal gas calculator, at atemperature of 400 R (222 K), the mol count would equal about 0.02 mols,which seems about right. At or slightly before TDC, the lower displacerexhaust valve will have closed. The lower displacer exhaust valve is acam-actuated poppet valve biased by a spring (not shown) towards closed,and is used to maintain a higher pressure within the lower displacerduring its intake stroke, as will be shown. Note that a small quantityof heated, pressurized H2 working fluid will remain in the lowerdisplacer following closure of the lower displacer exhaust valve at TDC.With lower displacer piston expansion past TDC, the pressure of theremnant gas in the lower displacer will immediately drop, eventuallymatching the 345 kPa (50 psi) pressure of the H2 at 400 R (222 K) in thelower displacer inlet manifold (not shown). The cold H2 lower displacerintake check valve (G) will thus be primed to overcome its light springbias and open, filling the lower displacer with H2 at 400 R (222 K) and50 psi (345 kPa). Note that, while the lower displacer cylinder volumeequals 4.86 cu in (0.08 L), the quantity of new H2 working fluid takeninto the lower displacer will be reduced by the amount of any remnantworking fluid in the lower displacer expanded back to 222 K and 50 psi.Assuming no remnant, maximum mol count per the ideal gas law calculatorwould approximate 0.015 moles of H2 per cycle. Per the CGL calculator,following a pure isochoric displacement of 0.08 L (4.86 cu in) of H2 atan initial temperature of 222 K (400 R) and an initial pressure of 345kPa (50 psi), for a final temperature of 400 K (720 R), internal energychange and heat would equal 0.0545 kJ (0.0516 BTU), and final pressurewould equal 622 kPa (90.2 psi).

Near TDC—Remnant Gas Recompression and Pressure Equalization:

Assuming source heat is added by isochoric combustion of H2 and oxygen,just previous to combustion, the previous charge of i.c.-heated H2+H2O(and/or H2O2 if H2O2 is used as the oxidizer) gaseous/vaporous mix willhave been exhausted from the expander cylinder (H). During the exhaust,the expander piston (I) moves from BDC towards TDC and thegaseous/vaporous mix will be exhausted past the expander exhaust valve(J) at approximately constant pressure and temperature, as will beshown. The expander exhaust valve is a cam-activated poppet valve with aspring bias towards closed (not shown), with any high pressuredifferential across the expander exhaust valve head thus sealing thevalve closed.

Just prior to TDC, the expander exhaust valve will be closed by thevalve's spring bias. This will result in the expander piston rapidlydriving any remnant working fluid captured in the expander to a higherpressure and temperature, eventually matching the pressure (80 psi (551kPa)) on the other side of the expander intake transfer valve (K), thusfreeing the expander intake transfer valve to pop towards open via areturn spring bias (not shown), and thus connecting the expander, theinternal volume of any external heater (L.), and the various connectingmanifolds between the transfer valve and the regenerator exhaust checkvalve. For insurance, a slight mechanical “bash” by the expander pistontop is arranged just before TDC to ensure the expander intake transfervalve does in fact open in a timely fashion just prior to i.c. H-in.

At or near TDC, isochoric i.c. will occur, for example by theinstantaneous connection, injection, and combustion of a sufficientquantity of pure O2 gas, for example at approximately 150 psi (1,034kPa). The combustion will instantly increase chamber pressure toapproximately 117.5 psi (810 kPa). Note that the back-flowing pressurewave will instantaneously seal the regenerator exhaust check valve (M),thus isolating the regenerator (B) from the pressure surge.

TDC to BDC—Expansions (Expander):

Thermal inputs other than isochoric and expansions other than adiabaticare clearly possible and will be discussed below. That includes what canbe termed “displacement heating and expansion” processes. In theinstance of a pure isochoric i.c. source heat addition and theinstantaneous combustion of H2 and O2 immediately following the chargingof the upper displacer cylinder, and assuming no H-in via an externalheater and direct injection of the upper displacer into the expander,the volume of working fluid prior to expansion will equal the upperdisplacer volume (4.86 cu in, 0.0796 L) and the volume of working fluidfollowing expansion will equal the expander volume (13.5 cu in, 0.221L). 1 g of H2 has a low heat value of 120 kJ or 120,000 J. Increasingthe internal energy change of the H2+O2 mix by 51.1 J thus requires thecombustion of 0.000426 g or 0.426 mg of H2. One mol of H2 has a mass of2.0158 g, and the H2 at TDC prior to combustion would equalapproximately 0.02 moles, or a mass of 40.32 mg. The mass of H2combusted will thus equal about 1% of the total mass of H2 in thecharge. The mass of O2 combusted per charge would be 8 times the mass ofthe H2 combusted, or about 3.3 mg of O2.

In one expansion iteration (FIG. 6 through FIG. 9 ). Assuming zerothermal input following TDC, the following upper displacer piston (N)and simultaneous expander piston movements past TDC and towards BDC willcreate an ideally adiabatic “displacement expansion”, with expanding gasbeing continuously moved from the lower displacer into the expanderduring the expander piston downstroke. Total expandable volume at TDCwill thus equal the combined volume of the upper displacer cylinder(0.079 L (4.86 cu in)), and the internal volume of the external heaterand manifolding (0.016 L (1 cu in)), or 0.095 L (5.85 cu in) (see “Knownand estimated volumes” above). Following combustion and still at TDC,temperature will vary throughout the mix, but ideal peak temperaturewill be held to approximately 1000 R (555 K) (see Point [3] on FIG. 19).

Note: Permissible peak temperature for an adiabatic expansion processmay permissibly be considerably higher than for an isobaric process,which may permissibly be considerably higher than for an isothermalprocess, since the average temperature over the expansion will be lower.An average working fluid temperature that is lower may permit use ofnon-lubricated bearing surfaces at a higher peak temperature. Theextents of this will need to be determined experimentally.

A. Isochoric Source Heat Input/Adiabatic Expansion

In the instance of a pure isochoric i.c. source heat addition via theinstantaneous combustion of H2 and O2 following the charging of theupper displacer cylinder, and assuming no H-in via an external heater, afollowing adiabatic/isentropic expansion can be made to take place (seePoint [4] on FIG. 19 ). The volume of working fluid prior to expansionwill equal the external heater (˜1 cu in, 0.0164 L), the external heaterconnecting manifolds (˜0.5 cu in, 0.008), and the upper displacer volume(4.86 cu in, 0.0796 L), or a total of 4.88 cu in (0.08 L). The volume ofthe working fluid following expansion will be equal to the externalheater (˜1 cu in, 0.0164 L), the various connecting manifolds (˜0.5 cuin, 0.008), and the expander volume (13.5 cu in, 0.221 L), or a total of13.65 cu in (0.224 L), or an expansion ratio of ˜2.8/1.

An adiabatic/isentropic expansion of 1 to 2.8 equals an expansion from˜3.4 cu ft to ˜9.5 cu ft (269 L). Per FIG. 19 , for a temperature dropfrom 1000 R to ˜720 R (400 K), pressure drops to ˜30 psi (207 kPa). Perthe gas law calculator, H2 mol count cycle per minute would equal 16.65moles at 1 atm (101 kPa) and 650 R (351 K) and a volume of 17.0 cu ft(488 L), which closely matches the 17 cu ft, and estimated mol count of16.9 moles of air plus clearance gases, in FIG. 18 . Recall also that amaximum of 0.015 moles of H2 at a temperature of 400 R (222 K) and apressure of 50 psi (345 kPa) is injected into the cycle via the lowerdisplacer. Finally, recall also that it was determined that FIG. 18 mayrepresent an engine undergoing multiple identical cycles.

Therefore, assuming that 16.65 moles are cycled per minute, a processthat cycles ˜0.015 moles can be said to be representing an engine with acycling speed of ˜1,130 cycles/revolutions per minute (rpm).

In FIG. 20 , an estimate of the pressure and volume information shown inFIG. 18 was estimated and inserted into a spreadsheet. The resultcreates the graphed curves shown in FIG. 20 . Total work out for theequivalent of 17 cu ft of H2 cycled over 1 minute is then calculated infoot-pounds/minute for the adiabatic expansion following isochoric (CV)process shown in the top chart, for the isobaric (CP) source heat inputprocess shown in the middle chart, and for the combined isothermal (CT)expansion/source heat input process shown in the bottom chart. Pressuresare entered in 2.5 psi differences between the estimated 117.5 psi peakpressure and the estimated 30 psi minimum pressure following expansion.

For the isochoric/adiabatic expansion process, total work out iscalculated in FIG. 20 to equal ˜46,270 ft lb/min, or 59.46 BTU/min or3,568 BTU/hr, or about 1.40 ho (1.046 kW, 3,764 kJ).

Note: 59.46 BTU/hr can be compared to FIG. 18 and the Btu output shownthere for the isochoric/adiabatic expansion. Assuming 17 cu ft flow perminute, or 0.015 moles of H2 per revolution at 1,130 rpm, since FIG. 18indicates Btu drops from ˜100 to ˜40 or a difference of ˜60 Btu/min or˜3,600 BTU/hr. Per the CGL calculator, for an adiabatic expansion of H2from (4.86 cu in, 0.0796 L) at an initial temperature of 555 K (1,000 R)and an initial pressure of 863 kPa (125 psi) to a final volume of 0.221L (13.5 cu in ( ), temperature would equal 370 K (666 R) (a little lowerthan estimated) and pressure would equal 207 kPa (30 psi). In addition,internal energy change and W-out would both equal 0.058 kJ. At 1,130 rpmover an hour, total H-in and W-out would equal 3,932 kJ (1.092 kWh,1.465 HP/hr).

B. Isobaric Source Heat Input/Adiabatic Expansion

In the isobaric/adiabatic SD CCVC engine, as in the isochoric/adiabaticSD CCVC engine, the start point for the isochoric waste heatregeneration process is BDC, where the H2 working fluid in the lowerdisplacer cylinder and the upper displacer volume (4.86 cu in, 0.0796 L)is assumed pressurized to about 3.5 atm (˜50 psi, 355 kPa) at atemperature of 400 R (222 K) (see Point [1] on FIG. 19 ). Assumingperfect isochoric regeneration of exhaust waste heat from the previousisobaric source heat addition and adiabatic expansion, per FIG. 19 ,isochoric waste heat regeneration will raise the approximately 0.015moles of H2 to about 790 R (439 K). (Note that this is a highertemperature at which source heat is added than for theisochoric/adiabatic SD CCVC engine mapped above. That is because: (1)All the proposed cycle models are (a) set to a given peak temperatureand (b) use the same compression process; and because, (2) as is shownin FIG. 19 , an isobaric thermal input inherently drives the adiabaticexhaust process to the right in comparison with an isochoric thermalinput, forcing the adiabatic line to connect with the volume limitingline of the isochoric regeneration process at a higher peaktemperature.) Per the CGL calculator, that higher input temperature fromthe isobaric/adiabatic SD CCVC engine exhaust will increase the pressurefollowing isochoric thermal regeneration to about 702 kPa (102 psi). InFIG. 19 , an isobaric source heat input/adiabatic expansion process isillustrated, beginning at the point where i.c. isobaric source H-in isbegun, that is, at TDC (see Point [5] on FIG. 19 ). In the instance of apure isobaric i.c. source heat addition and the instantaneouscommencement of combustion of H2 and O2 immediately following thecharging of the upper displacer cylinder, and assuming direct injectionof the upper displacer into the expander, the volume of working fluidprior to expansion will equal the upper displacer volume (4.86 cu in,0.0796 L) and the volume of working fluid following expansion will equalthe expander volume (13.5 cu in, 0.221 L). The isobaric expansion thusbegins at the pressure, temperature, and volume of the H2 working fluidin the upper displacer at TDC, or ˜702 kPa (102 psi), 790 R (439 K), and(4.86 cu in, 0.0796 L).

Per the CGL calculator, the isobaric expansion completes at ˜702 kPa(102 psi), 1000 R (555 K), and ˜0.101 L (6.16 cu in), with H-in equal to0.0511 kJ (0.0480 BTU) and W-out equal to 0.0148 kJ (0.0140 BTU). (seePoint [6] on FIG. 19 ) At 1,130 rpm over an hour, that equals 1,003 kJ(0.279 kWh, 0.374 HP). Total H-in at 1,130 rpm over an hour equals 3465kJ (3,284 BTU).

Immediately following isobaric expansion, adiabatic expansion continuesuntil volume equals 0.221 L (13.5 cu in) (see Point [7] on FIG. 19 ).Per the CGL calculator, the adiabatic expansion of H2 from 0.101 L (6.16cu in), 555 K (1,000 R) and 702 kPa (102 psi) to a volume of 0.221 L(13.5 cu in) results in a final pressure of 234 kPa (34.9 psi) and afinal temperature of 404 K (727 R). The internal energy change and W-outequal 0.0475 kJ (0.045 BTU). At 1,130 rpm over an hour, that equals3,221 kJ (0.895 kWh, 1.20 HP). Total work out would thus equal 4224 kJ(1.17 kW, 1.57 HP).

C. Isothermal Source Heat Input/Expansion

Perhaps the most intriguing aspect of the proposed SP-OCVC and MCSP-OCVCdesigns is the ability to make possible a highly efficient isothermalexpansion process. In FIG. 19 , 1000 R (555 K) was “set” as the peaktemperature that a non-lubricated engine can sustain. Futureexperimentation will of course determine the true sustainable peaktemperature. But even at 1000 R, the ability to regenerate heat from theexhaust into the working fluid following expansion will have a profoundimpact on developing highly efficient heat engines.

In FIG. 19 , an isothermal expansion from 0.0796 L (4.86 cu in) isillustrated (see Point [3] on FIG. 19 ). Note that this temperature isachieved by the theoretical 100% thermal regeneration of the 555 K(1,000 R) exhaust. Isothermal expansion, which is assumed to result fromcombustion of H2 and injected O2 internally, will then continue to (13.5cu in, 0.221 L). Per the CGL calculator, an isochoric heat input into0.0796 L (4.86 cu in) of H2 at an initial temperature of 222 K (400 R)and pressure of 345 kPa (50 psi) to an final temperature of 555 K willresult in a final pressure of 863 kPa (125 psi). Per the CGL calculator,an isothermal expansion from 0.0796 L (4.86 cu in) at a temperature of555 K to a final volume of 0.221 L (13.5 cu in) would result in a finalpressure of 311 kPa (45.1 psi) and H-in and total W-out would equal0.070 kJ. Assuming 1,130 rpm, W-out and H-in over 1 hour for theisothermal expansion process equals 4,746 kJ (1.32 kWh, 1.77 HP/hour).

D. Displacement Expansion

A typical stirling engine uses displacement heating, displacementcooling, displacement compression, and displacement expansion. Theproposed SD-CCVC engine can be seen to use the same kind of processes,albeit in separate and discrete segments. Testing of the CCVC prototypegave some data indicating the probable effect of displacement heating inthat design.

From General Report on the Close Cycle Valved Cell (CCVC) heat enginetest program to the California State Energy Innovation Small Grant(EISG) program administrator, July, 2005 (reference to colors will beclarified below):

“For all tests, a driver motor drove the prototype. As in previoustests, by rotating the engine with a driver motor, pressure transducerswere able to “map” some of the internal pressure fluctuations of theengine . . . . These maps show the basic pressure fluctuations occurringwithin the engine at various temperature inputs and rpm's. Note that thepressure lines are formed by pressure readings captured from thetransducers 5000 times per second. Rpm is thus determined by measuringthe number of readings per complete cycle (for example, a complete cyclethat occurs in 2500 readings indicates a cycle operating at 120 rpm).The red line represents the pressure fluctuations occurring in theengine's heater manifold (the upper displacer/expander “side” of theengine) and the blue line represents the pressure fluctuations occurringin the engine's cooler manifold (the compressor/cooler displacer “side”of the engine). The lines running from top to bottom representapproximate TDC and BDC in the engine.

It is important to keep in mind that the physical makeup of upperdisplacer and compressor “sides” of the engine changes during an enginecycle, as various valves connect or disconnect various volumes. As aresult, the expander will be connected to the upper displacer chamberduring expansion and to the compressor during exhaust, while the lowerdisplacer chamber will be connected to the compressor during compressionand to the upper displacer chamber during exhaust. This means that thepressure transducers only observed part of what was happening within theengine. The rest must be deduced, as will be shown.”

FIG. 21 is from Chart 14 of the July, 2005 EISG report. (NOTE: This is agreyscale conversion of this data which de-colorizes it. Only the “red”and “dark blue” lines are actual transducer data. The jagged linestarting at the top left and ending second from the top right is the“red” transducer data line and the jagged line starting at the bottomleft and ending at the bottom right is the “blue” transducer data line.)FIG. 21 indicates that the point was reached within the CCVC prototypewhere W-in equaled W-out, assuming zero friction, pumping, and thermallosses. It also suggests that the expansion process was not adiabatic,to wit:

On the left top quadrant, the “red” descending line indicates theobserved expansion process, up until the transfer valve closed at about135 psi. (The red re-ascending line moving to BDC represents there-compression by the hot displacer into the dead space between theexpansion cylinder transfer (intake) valve and the cold displacerexhaust check valve, which includes the heater internal volume, therecuperator internal volume, and the various connecting manifoldvolumes. Note that, as the cold displacer displaces into and through therecuperator at constant volume, the pressure increases as TDC isapproached, as predicted.)

On the right top quadrant, the “dark blue” line descending from BDC andabout 170 psi indicates an nearly adiabatic compressor gas re-expansionprocess, up until the exhaust check valve opens at about 115 psi. (The“dark blue” descending line from that point indicates the constantvolume displacement of hot gas out of the expander, into and through theexpander inlet side of the recuperator, through the cooler, and into thecompressor. Note that, as the hot gas displaces into and through therecuperator and cooler at constant volume, the pressure decreases as TDCis approached, as predicted.)

It is visually obvious that the displacement expansion process is notadiabatic, but that heat is being added to the fluid displacing from thedisplacer, through the heater, and into the expander. Therefore, whileit is not possible with just this small amount of data to predictexactly how a displacement expansion curve would track, it is clear thatthe peak temperature and pressure are quite a bit higher than the finalexpansion temperature and pressure. That suggests that, as with theisochoric/adiabatic expansion process (and the isobaric/adiabaticexpansion process), the peak temperature can be elevated substantiallyabove the overall temperature limit required to operate withoutlubricant.

It can therefore be assumed that peak temperature for an SD-CCVCdisplacement expansion engine as modeled above could approximate that ofan isochoric engine as modeled above and stay within this limit,although more work would be generated and more input heat would berequired with a displacement expansion process than with anisochoric/adiabatic expansion process. Actual determination of thermalefficiency and power output for a pure SD-CCVC displacement expansionengine would thus approximate that of an SD-CCVC isochoric/adiabaticexpansion engine, but can't be accurately determined until actualtesting is carried out.

One further interesting takeaway from FIG. 21 : Consideration was notgiven at the time to the potential of reducing the sink temperature, asby running the CCVC engine within a PSR on the lunar surface. If thesink temperature were reduced by 200-300 K, that would be effectivelythe same as increasing the peak temperature by the same amount. Pressuredifferential would be accordingly raised, and expansion would proceedmore deeply than seen in FIG. 21 .

BDC to TDC—Expander to SD Cylinder:

At BDC (on FIG. 19 , see Point [4] for the isochoric/adiabaticexpansion, see Point [7] for the isobaric/adiabatic expansion, and seePoint [8] for the isothermal/adiabatic expansion), closing the upperdisplacer exhaust/expander intake/transfer valve would decrease thevolume available for the following exhaust processes to 13.5 cu in(0.221 L). Per FIG. 20 , the pressure following expansion equals ˜27.5psi for the isochoric/adiabatic process (see Point [3] on FIG. 19 ),˜32.5 psi for the isobaric/adiabatic process (see Point [6] on FIG. 19), and ˜40 psi for the isothermal process (see Point [8] on FIG. 19 ).Assuming the exhaust valve completely closes with ⅛″ of travel, sincearea equals 4.91 sq in across the piston head, volume equals 0.61 cu in(0.010 L). Per FIG. 19 , temperature equals 720 R (400 K) and pressureequals ˜30 psi (207 kPa). Per the ideal gas calculator, at a temperatureof 720 R (400 K), a volume of 13.5 cu in (0.221 L), and a pressure of207 kPa, molal mix equals 0.0006 moles. If 0.015 moles transfer, then atmaximum expansion, the expander holds 0.0156 moles of working fluid mix.Per the ideal gas calculator, pressure equals 241 kPa (35 psi). Theexhaust process discussed below will be similar for all three expansionprocesses, only differing in the pressure and temperature of theexhausting working fluid mix, but will be assumed for the purposes ofcalculation to be isobaric at 30 psi (207 kPa). (Note: The pressureresults following expansion per the CGL calculator are also different.However, for purposes of generating a 1st order calculation and definingthe processes, 30 psi (207 kPa) is deemed acceptable.

As stated above, an SD cylinder enables use of a stirling cycle-typeregenerator. At TDC, the SD cylinder (O), via an insulated expanderexhaust manifold (not shown), will have received a full charge from theexpander cylinder (H) at approximately constant temperature, pressure,and volume, with a small volumetric reduction relative to that of theexpander cylinder resulting from the 1st stage compressor-to-SD cylinderconnecting rod (T).

Thus, for the isochoric/adiabatic expansion:

Expander cylinder volume equals 13.5 cu in (0.221 L) and expander pistonarea equals 4.91 sq in. Force of exhaust thus equals 147 pounds over2.75 inches or 0.227 feet or 33.4 ft lb. At 1,130 rpm, that equals37,784 ft lb/minute or 48.5 BTU/minute or 2,913 BTU/hr, or 1.14 HP/hr(0.854 kWh, 3.073 kJ/hr).

The SD cylinder internal volume is slightly less than the volume of theexpander, equaling 13.39 cu in (0.219 L). SD piston (S) area thus equals4.87 sq in. The intake W-out thus equals 146 pounds over 0.227 feet or33.2 ft lb. At 1,130 rpm, that equals 34,353 ft lb/minute, or 1.04 HP/hr(0.776 kWh, 2,792 kJ). The expander exhaust-to-SD cylinder displacementprocess therefore requires approximately 0.01 HP/hr of W-in. That is anegligible amount, and for that reason calculations for the isobaric andisothermal expansion processes will also be assumed to be negligible.

Exiting the expander, the exhaust passes through the expander exhaustvalve, into the expander-to-SD manifold (not shown), past the SD inletcheck valve (AE), past the SD inlet actuated valve (Q), and into the SDcylinder. The SD inlet check valve in this instance is a simple poppetvalve lightly biased towards closed. That is, a small pressuredifferential across the valve head will open the valve and allow flowfrom the expander exhaust manifold to pass, but any higher pressure onthe SD cylinder side of the valve head will cause the valve to firmlyshut.

The SD cylinder inlet actuated valve is a poppet valve (mechanicallyclosed by a crankshaft-mounted camshaft/pushrod/rocker arm-actuatedassembly) that is biased towards open, but blocking gas flow in theopposite direction to the SD cylinder inlet check valve. That is, whenthe SD cylinder inlet actuated valve is closed, working fluid mix fromthe expander exhaust manifold cannot go into the SD cylinder.

Note: When pressure in the SD cylinder increases over the pressureacross the SD cylinder inlet actuated valve head, (which occurs once percycle), some leakage can occur. That is the reason for the SD cylinderinlet check valve's presence; that is, the SD cylinder inlet check valvewill not allow back-pressure flow into the expander exhaust manifold dueto leakage past the SD cylinder inlet actuated valve head.

The SD cylinder inlet actuated valve will manually connect anddisconnect the SD cylinder and the expander-to-SD manifold when thepressure in the SD cylinder approximately equals the pressure andtemperature of the gas exhausting from the expander, which occurs at twodistinct points per cycle, as will be shown below. As a result, the SDcylinder inlet check valve will maintain the pressure in theexpander-to-SD manifold at times when the SD cylinder pressure is higherthan the expander exhaust manifold pressure, and the SD cylinder inletactuated valve will maintain the pressure in the expander-to-SD-manifoldat times when the SD cylinder pressure is lower than the exhaustmanifold pressure. The SD cylinder inlet actuated valve opens (that is,is actuated against the bias) slightly before BDC, and the SD cylinderinlet actuated valve closure occurs at or slightly before TDC, as willbe shown below.

TDC to BDC—SD Cylinder to 1st Stage Compressor Via Regenerator.

Starting at TDC (on FIG. 19 , see Point [4] for the isochoric/adiabaticexpansion, see Point [7] for the isobaric/adiabatic expansion, and seePoint [8] for the isothermal/adiabatic expansion), the SD piston, withapproximately a full charge of working fluid mix, will travel downward.At TDC, the SD cylinder exhaust poppet-type check valve (R) is beingheld closed by pressure differential across the valve head. Recall that,by the time TDC has been reached, the working fluid in the lowerdisplacer, regenerator, and upper displacer has reached the maximumpressure just prior to i.c. initiation. SD piston travel from TDC thusbegins compressing the captured working fluid mix within the SDcylinder, raising its pressure. Pressure differential across the SDcylinder exhaust check valve (R) will keep that valve from opening,while the SD cylinder intake check valve (AE) will keep any pressurizingworking fluid mix within the SD cylinder from entering the expanderexhaust manifold.

Simultaneously at TDC, the 1st stage compressor piston (W) will beginexpanding working fluid out of the regenerator via the 1st stagecompressor cylinder (V) intake transfer valve (U), since the 1st stagecompressor intake transfer valve will have previously opened, as will beshown below. This 1st stage compressor piston expansion will thussimultaneously drop pressure in the 1st stage compressor cylinder andthe connected regenerator while the pressure is being raised in the SDcylinder. Consequently, pressure will quickly equalize across the SDcylinder exhaust check valve, allowing a flow of hot working fluid mixfrom the SD cylinder to pass through the SD cylinder exhaust checkvalve, past the regenerator, past the previously opened 1st stagecompressor intake transfer valve (as will be shown below), and into the1st stage compressor cylinder.

Note: When the SD cylinder exhaust check valve opens as a result of thepressure equalization, the SD cylinder's adiabatic pressurizationprocess and the 1st stage compressor cylinder adiabatic depressurizationprocess will be instantly convert to an isochoric displacement process.Note that the volume of the SD cylinder and the 1st stage compressorboth exactly equal 13.39 cu in (0.219 L). As isochoric displacement isinitiated, heat will be removed at constant volume from the workingfluid mix exiting the SD cylinder by the regenerator, causing theworking fluid mix pressure and temperature to instantly begin to drop.Note that this isochoric displacement will begin when the pressurewithin the SD cylinder drops approximately to that of the working fluidmix in the regenerator+regenerator manifold+1st stage compressor.

In the regenerator at TDC, pressure equals 80 psi (551 kPa) and volumeequals 1.5 cu in (0.0246 L). Per FIG. 18 , the ideal temperature of theworking fluid entrapped in the regenerator would equal ˜720 R (400 K)(that is, the temperature of the gas exhausting from the expander).However, the average temperature of the entrapped working fluid acrossthe regenerator at 80 psi would equal the average difference between thepeak temperature and the lower displacer inlet temperature (400 R, 222K), which would equal 560 R (311 K). Recall that the estimatedregenerator and manifold internal volume would equal 1.5 cu in (0.0246L). From the ideal gas calculator, mol count in the regenerator andmanifolding at 311 K (560 R) and 551 kPa (80 psi) will thus be estimatedat 0.0052 moles. Assuming an expansion of densified working fluid at thelow end temperature of ˜400 R (222 K), very little expansion of the 1ststage compressor cylinder and compression of the SD cylinder will berequired to equalize pressure across the SD cylinder exhaust check valveto something less than 80 psi and greater than 30 psi. Moving past TDC,the pressure within the SD cylinder will increase adiabatically over thebeginning pressure of 30 psi. At the same time, the pressure within the1st stage compressor will drop adiabatically under 80 psi. Assuming thecombined 1st stage compressor and regenerator space volume doubles from1.5 cu in (0.0246) to 3 cu in (0.049 L), all in the low temperature sideof the regenerator, and assuming an average temperature drop in thecombined 1st stage compressor and regenerator space volume to ˜500 R(278 K) at the end of the pressure equalization process, and furtherassuming 0.0052 moles of gas, then per the ideal gas calculator,pressure would drop from 551 kPa (80 psi) to 241 kPa (35 psi), and theSD cylinder exhaust check valve would have opened. Per the ideal gascalculator, 0.015 moles of H2 at 400 K (720 R) in the SD cylinder with avolume of 0.219 L (13.39 cu in) would equal a pressure of 228 kPa (33.1psi), or slightly higher than earlier estimates. Per the CGL calculator,increasing pressure from 228 kPa (33.1 psi) to 241 kPa (35 psi) willdecrease volume to 0.211 L (12.8 cu in). That is generally in the ballpark of what would be required to equalize pressure across the SDcylinder exhaust check valve. Over a 1st stage compressor piston area of4.87 sq in, and a cylinder volume increase of an estimated 1 cu in, or achange in volume of 1/13.39th or 7.47%, that would equal a traveldistance of ˜0.205″. Note that the same ˜0.205″ would be traveled by theSD cylinder as it compresses the working fluid mix entrapped within it.This would amount to a negligible amount of W-in by the SD cylinderbefore pressure is equalized across the SD cylinder exhaust check valve.Also, W-out by the 1st stage compressor is expected to more than balanceW-in to the SD cylinder.

As the 1st stage compressor approaches BDC (for all SD CCVC models, seePoint [9] on FIG. 19 ) during isochoric displacement, pressure andtemperature will have dropped. Per FIG. 19 , pressure in the SDcylinder, the regenerator, and the 1st stage compressor would drop toabout 25 psi (172 kPa), assuming a complete displacement. Per the CGLcalculator, isochoric expansion starts at 241 kPa (35 psi) with the vastmajority of gas in the SD cylinder at a temperature of ˜400 K (720 R)and a volume of 219 L (13.39) cu in (not inclusive of the regeneratorand manifolding). Assuming a final temperature within the 1st stagecompressor of 234 K (421 R), final pressure would equal 141 kPa (20psi). Pressure is expected to be slightly higher, however, since thereis expected to be an additional 0.001 moles of H2 within the 1st stagecompressor due to clearance gas carrying over from the previous stroke.(Note: The SD cylinder actuated inlet valve (Q), being at this pointclosed, prevents a “blowdown” in gas from the expander-to-SD manifold(not shown) into the SD cylinder due to the drop in pressure below 30psi.) However, it is essential that pressure be equalized across thelower displacer exhaust valve (E) before that valve opens at or nearBDC. Recall that pressure on the displacer side of the lower displacerexhaust valve is at ˜50 psi. In order to increase the pressure on theregenerator side of the lower displacer exhaust valve, the 1st stagecompressor intake transfer valve (U) is closed early, as BDC isapproached. This will immediately cause the SD piston to recompress anyworking fluid mix now captured between the SD piston, the lowerdisplacer exhaust valve, and the 1st stage compressor intake transfervalve. In other words, the 1st stage compressor intake transfer valvewill be designed to close at a point that allows movement of the SDpiston to BDC to reach the pressure found in the lower displacercylinder, or ˜50 psi.

Earlier, it was calculated that, in the regenerator and its manifolds atTDC, mol count equals 0.0052 moles at an average temperature across theregenerator of ˜311 K and 80 psi. As noted above, per FIG. 19 , pressurein the SD cylinder, the regenerator, and the 1st stage compressor woulddrop to about 25 psi (172 kPa), assuming a complete displacement.Assuming a pressure nearing full displacement of about 25 psi (172 kPa),at an average temperature of 311 K (per above) in the regenerator andits manifolding (equal to 1.5 cu in or 0.0246), mol count in theregenerator and its manifolding, per the ideal gas law calculator, willdrop to about 0.00164 moles. Following closure of the 1st stagecompressor transfer valve at an average temperature of 311 K and 50 psi(345 kPa), per the ideal gas law calculator, mole count in theregenerator+manifolds or 1.5 cu in (0.0246) would equal 0.00328 moles.Since 0.00328 moles are required, approximately 0.00164 moles mustremain in the SD cylinder following displacement and prior to reachingBDC, or just sufficient “dead space” moles to allow recompression backto ˜50 psi at BDC, thus allowing the lower displacer exhaust valve (E)to open without a pressure difference.

Per FIG. 19 , the temperature of the entrapped working fluid in the SDcylinder just following 1st stage compressor transfer valve closure atabout 25 psi (172 kPa) is about 525 R (292 K). Per the ideal gas lawcalculator, with a mol count of 0.00164, that would equal a volume of0.00023 L. Since the total volume of both the SD cylinder and the 1ststage compressor cylinder equals 0.219 L (13.39 cu in), total volume ofthe 1st stage compressor cylinder at the time of the 1st stagecompressor intake transfer valve is essentially unchanged at 0.219 L(13.39 cu in).

At the instant the 1st stage compressor transfer valve closes, assuming0.015 moles pass through to the 1st stage compressor, and 0.001 molesremained in the 1st stage compressor cylinder clearance space at TDC,approximately 0.016 moles remain in the 1st stage compressor cylinder.From above, pressure would have dropped to about 25 psi (172 kPa), andvolume would equal 13.39 cu in (0.219 L). Per the ideal gas calculator,temperature would thus equal ˜234 K (421 R), which is a close estimateof the temperature at the point in FIG. 19 (for all SD CCVC models, seePoint [9] on FIG. 19 ).

Since the displacement of the working fluid mix from the SD cylinder andinto the 1st stage compressor occurred at constant volume, only anegligible amount of net W-in or W-out occurs.

Mechanically, as BDC is approached, the SD cylinder actuated inlet valveactuator (not labeled) will allow the SD cylinder actuated inlet valve(Q) to open, since it is biased towards open. Note that a small amountof pressurization into the space between the SD cylinder actuated inletvalve and the SD cylinder intake check valve (AE) will then take place,increasing to about 50 psi. Further note that the SD cylinder intakecheck valve will not permit back flow into the expander-to-SD manifold(not shown). As shown in FIGS. 11 thru 17, the SD cylinder actuatedinlet valve actuator is an arm physically touching the stem of the 1ststage compressor actuated exhaust valve (X). That SD cylinder actuatedinlet valve actuator, which is physically connected to the 1st stagecompressor actuated exhaust valve, closes the SD cylinder actuated inletvalve when the 1st stage compressor actuated exhaust valve closesslightly before TDC, as will be shown. Since the 1st stage compressoractuated exhaust valve opens just prior to BDC, the SD cylinder actuatedinlet valve, which is biased towards open, will also be allowed to openjust prior to BDC, assuming pressure differential allows it. However,note that, prior to the closing of the 1st stage compressor transfervalve, there will be a pressure differential across the SD cylinderactuated inlet valve head as a result of the lower pressure (˜25 psi) inthe SD cylinder versus in the manifold space (˜30 psi) between the SDcylinder actuated inlet valve and SD cylinder inlet check valve (AE). Asa result, the SD cylinder actuated intake valve itself will remainclosed until pressure builds (<30 psi) in the SD cylinder sufficient totrigger the bias towards open, at which point the valve willautomatically open. (Note: It is possible that a small amount ofpressurized working fluid will escape past the SD cylinder check intakevalve, but it should be minimal.)

With further movement of the SD piston to BDC, the working fluid mixremnant (within the SD cylinder, the regenerator and its manifolds, andthe manifold connecting the SD cylinder intake check valve) will climbin pressure above the pressure acting to open the SD cylinder inletcheck valve (lightly biased towards closed), closing that valve, ifopen. As pressure in the cylinder grows above ˜30 psi to ˜50 psi, thepressure in the expander exhaust manifold will be kept constant (at ˜30psi) by virtue of the SD cylinder's inlet check valve.

As a result of recompression of remnant gas in the SD cylinder as the SDpiston moves to BDC, the pressure in the SD cylinder will reachapproximately the pressure in the lower displacer cylinder (˜50 psi),allowing the lower displacer actuated exhaust valve to be mechanicallyopened without incurring a pressure drop.

(Note: The push rod that operates the lower displacer actuated exhaustvalve (not labeled) is “shortened” (that is, some cam lift is allowed tooccur before the push rod contacts the lower displacer actuated exhaustvalve) to allow the compression by the SD piston to largely go tocompletion at BDC before the lower displacer cylinder exhaust valve isactuated towards open. It is considered obvious that other means forsuccessfully operating the lower displacer actuated exhaust valve arepossible.

BDC to TDC—1st Stage Compressor Compression:

Note: The 1st stage compression is assumed to use purely adiabaticcompression coupled with a purely isobaric exhaust processes. However,to the degree that the adiabatic process actually approaches isothermal,less work will be required for a given pressure to be achieved. It isexpected that the walls and internal piston of the compressor will bechilled to the “ambient” temperature of 400 R, and thus some degree ofheat transfer out of the compressing working fluid mix is expectedduring compression.

Per the CGL calculator, for an adiabatic compression of 0.219 L (13.39cu in) of H2 from an initial temperature of ˜234 K (421 R) and aninitial pressure of 25 psi (172 kPa), for a final pressure of 207 kPa(30 psi), final temperature would equal 247 K (445 R) and would equal avolume of 0.192 L (11.7 cu in) (for all SD CCVC models, see Point [10]on FIG. 19 ). In addition, internal energy change and required W-inwould equal 0.0051 kJ. Assuming 1,130 rpm, W-in per hour for thecompression process equals 346 kJ (0.096 kWh, 0.13 HP/hour).

In addition, as noted above, 0.001 moles of the 0.016 moles in the 1ststage compressor cylinder is “reserved” for re-pressurizing the STREPregenerator (B), meaning total exhausted volume is equal to 15/16ths of0.192 L or 0.180 L (11 cu in). Further, the temperature of the 0.015 molof exhausted charge, which equals 247 K (445 R), will be cooled to 222 Kin the following isobaric cooler, reducing the exhausted volume, per theideal gas calculator, to 0.134 L (8.18 cu in)

An isobaric exhaust process at −30 psi proceeds from 0.180 L (11 cu in)until slightly before TDC (for all SD CCVC models, see Point [11] onFIG. 19 ), when the 1st stage compressor actuated exhaust valve isclosed. A small amount of remnant working fluid is thus captured withinthe 1st stage compressor, allowing that remnant fluid to be recompressedto the pressure on the regenerator side of the 1st stage compressortransfer valve (˜80 psi). The remnant fluid is assumed to equal 0.001moles, thus leaving 0.015 moles to be exhausted per stroke of the 1ststage compressor.

Note that the 1st stage compressor intake transfer valve, similarly tothe expander intake transfer valve, is opened by a combination ofpressure differential equalization and spring bias, then mechanicallyclosed, then kept closed by pressure differential. Also similarly, aslight mechanical “bash” by the 1st stage compressor piston top may bearranged just before TDC to ensure the 1st stage compressor intaketransfer valve opens in a timely fashion.

Total volume of the 1st stage compressor piston equals 0.219 L (13.39 cuin). Following adiabatic compression, 0.180 L (11 cu in) remain, or82.2% of total piston travel. Total piston travel is 2.75″, therefore,total travel at constant pressure equals 2.26″. Exhausting the workingfluid mix at 30 psi over a 4.91 sq in area equals 147.3 lb force.Traveling along a stroke of 2.26″ or 0.188′ creates a W-in of 27.7 ft lbper stroke. Total 1st stage compressor isobaric W-in at 1,130 rpm equals31,292 ft lb/minute or 0.948 HP/hr (0.71 kWh, 2,545 kJ). Total work infor the 1st stage compressor equals 346 kJ (0.096 kWh, 0.129 HP/hr) ofadiabatic compression plus 2,545 kJ (0.707 kWh, 0.948 HP/hr) of isobaricexhaust, or 2891 kJ (0.803 kWh, 1.07 HP/hr).

BDC to TDC—2nd Stage Compressor Isobaric Intake:

The 2nd stage compressor (A, AA) is also assumed in this analysis to usepurely adiabatic compression, in this case coupled with both an isobaricintake and an isobaric exhaust process. However, to the degree that theadiabatic compression process approaches isothermal, as by cooling ofthe cylinder and piston walls, less work will be required for a givenpressure to be achieved.

The 2nd stage compressor cylinder receives its charge as the 1st stagecompressor piston moves from BDC (for all SD CCVC models, see Point [9]on FIG. 19 ) to TDC (for all SD CCVC models, see Point [11] on FIG. 19), said first stage compressor piston simultaneously compressing andexhausting the H2 cooled by the regenerator through the 1st stagecompressor actuated exhaust valve (X), through the 1st stage compressorexhaust check valve (Y), through the 1st stage compressor exhaust cooler(Z), through the 2nd stage compressor intake check valve (AB) and intothe compressor cylinder. It therefore has a full charge at TDC at apressure of −30 psi and a temperature of ˜400 R (222 K).

Importantly, the working fluid mix, upon exiting the 1st stagecompressor exhaust cooler system, or some early portion of said system,can be seen to be below the temperature at which any H2O in the mix willhave liquified and condensed out of the H2 working fluid. It is a simplematter to then completely separate the H2O liquid and the H2 workingfluid. It is also a simple matter to add at some point sufficient new H2working fluid at ˜400 R (222 K) and 30 psi to achieve the ˜0.015 molesof H2 required to continue the overall cycle. One potential H2O removalsite is indicated in FIGS. 15 and 17 . The 2nd stage compressor volumeand piston area is formed by the space between the 2.5″ dia, 4.91 sq in(31.7 cm2) area, 13.5 cu in (0.221 L) volume of the 2nd stage compressorcylinder (A) and the 1.6″ (4.06 cm) dia, 2 sq in (12.9 cm2) area, 5.5 cuin (0.09 L) volume of the 2nd stage compressor piston connecting tube(AA). That equals a net volume of 8 cu in (0.131 L) and a net pistonarea of 2.91 sq in.

Per the ideal gas calculator, at 222 K, 30 psi (206.84 kPa), and 0.015moles, volume will equal 0.134 liters (8.17 cu in) The 2nd stagecompressor will thus closely match the 8.18 cu in 30 psi constantpressure output from the 1st stage compressor. (Note: the 1st stagecompressor exhaust cooler ensures that the output remains at 222 K (400R).) Note that the 2nd stage compressor cylinder piston (F) is on thelower side of the lower displacer-and-2nd stage compressor piston, whilethe lower displacer cylinder piston is on the upper side of the lowerdisplacer-and-2nd stage compressor piston. Total piston travel of the2nd stage compressor piston equals 2.75″, therefore, total travel atconstant pressure equals 2.75″ or 0.229′. Receiving the working fluidmix at 30 psi over a 2.91 sq in area equals 87.3 lb force. Travelingalong a stroke of 0.23′ creates a W-out of 20.0 ft lb per stroke. Total1st stage compressor isobaric W-out at 1,130 rpm equals 22,600 ftlb/minute or 0.685 HP/hr (0.510 kWh, 1,840 kJ).

TDC to BDC—2nd Stage Compression and Isobaric Exhaust:

Per the CGL calculator, an adiabatic compression of 8.17 cu in (0.134 L)of H2 from an initial temperature of 400 R (222 K) and an initialpressure of 30 psi (207 kPa) (for all SD CCVC models, see Point [11] onFIG. 19 ), to a final pressure of 345 kPa (50 psi) and a finaltemperature of 257 K (463 R), would equal a volume of 0.093 L (5.68 cuin) (for all SD CCVC models, see Point [12] on FIG. 19 ). In addition,internal energy change and required W-in would equal 0.0108 kJ. Assuming1,130 rpm, W-in per hour for the compression process equals 732 kJ (0.20kWh, 0.27 HP/hour).

Exhaust W-out can be calculated by MEP over area over length ofexpansion. The MEP equals 50 psi. Area equals 2.91 sq in. Force equals146 lb over an expansion of 0.229′. Force thus equals 33.32 ft lb. At1,130 rpm, that equals 37,651 ft lb/min, or 1.141 HP/hr (0.851 kWh,3,063 kJ). Total W-in of 2nd adiabatic expansion and isobaric exhaustthus equals 1,223 kJ (0.34 kWh, 0.46 HP/hr) (for all SD CCVC models, seePoint [1] on FIG. 19 ). Further, the temperature of the 0.015 mol ofexhausted charge, which equals 257 K (463 R), will be cooled to 222 K inthe following isobaric cooler, reducing the exhausted volume, per theCGL calculator, to 0.080 L (4.86 cu in) (for all SD CCVC models, seePoint [1] on FIG. 19 ).

BDC to TDC—2nd Stage Compressor to Lower Displacer:

From earlier, the upper and lower displacer cylinder volumes minus thedisplacer piston connecting tube volumes equals 4.86 cu in (0.0796 L)each. Since the stroke equals 2.75″, piston area equals 1.77″. Withintake pressure equal to 50 psi, force equals 88.4 pounds. over 0.229′,W-out per stroke equals 20.2 ft lb. Over 1,130 rpm and 60 minutes, totalft lb/hr equals 1,371,951 ft lb, or 1,860 kJ (0.52 kWh, 0.693 HP/hr)(for all SD CCVC models, see Point [1] on FIG. 19 ).

Typical SD-CCVC Cycle Net Work and Thermal Efficiency:

-   -   SD-CCVC Isochoric/Adiabatic Cycle Expansion W-Out:        -   1.465 HP/hr (1.092 kWh, 3,932 kJ)    -   SD-CCVC Isobaric/Adiabatic Cycle Expansion W-Out:        -   1.89 HP/hr (1.41 kWh, 5,071 kJ)    -   SD-CCVC Isothermal Cycle Expansion W-Out:        -   1.85 HP/hr (1.38 kWh, 4,970 kJ)    -   1st Stage Compressor Compression W-In:        -   0.129 HP/hr (0.096 kWh, 346 kJ)    -   1st Stage Compressor Exhaust W-In:        -   0.948 HP/hr (0.707 kWh, 2,545 kJ)    -   2nd Stage Compressor Expansion W-Out:        -   0.685 HP/hr (0.510 kWh, 1,840 kJ)    -   2nd stage compressor compression W-in:        -   0.27 HP/hr (0.20 kWh, 732 kJ)    -   2nd Stage Compressor Exhaust W-In:        -   1.141 HP/hr (0.851 kWh, 3063 kJ)    -   Lower Displacer Expansion W-Out:        -   0.693 HP/hr (0.52 kWh, 1,860 kJ)    -   Total 1st and 2nd stage compressors W-in, all cycles:        -   1.11 HP/hr (0.828 kWh, 2,979 kJ)    -   Total W-Out, SD-CCVC Isochoric/Adiabatic Cycle:        -   1.465 HP/hr (1.092 kWh, 3,932 kJ)    -   Net W-Out, SD-CCVC Isochoric/Adiabatic Cycle:        -   0.355 HP/hr (0.265 kWh, 953 kJ)    -   Thermal Efficiency, SD-CCVC Isochoric/Adiabatic Cycle:        -   Thermal source isochoric input: 3,726 BTU (3,932 kJ)        -   Thermal efficiency: 24.2%    -   Total W-Out, SD-CCVC Isobaric/Adiabatic Cycle:        -   1.57 HP/hr (1.17 kWh, 4,204 kJ).    -   Net W-Out, SD-CCVC Isobaric/Adiabatic Cycle:        -   0.464 HP/hr (0.346 kWh, 1,245 kJ)    -   Thermal Efficiency, SD-CCVC Isobaric/Adiabatic Cycle:        -   Thermal source isobaric input: 3,284 BTU (3465 kJ)        -   Thermal efficiency: 35.9%    -   Total W-Out, SD-CCVC Isothermal Expansion Cycle:        -   1.77 HP/hr (1.32 kWh, 4,746 kJ)    -   Net W-Out, SD-CCVC Isothermal Expansion Cycle:        -   0.65 HP/hr (0.49 kWh, 1,767 kJ)    -   Thermal Efficiency, SD-CCVC Isothermal Cycle:        -   Thermal source isothermal input: 4,498 BTU (4,746 kJ)        -   Thermal efficiency: 37.2%

Typical SD-CCVC Cycle H2 and O2 Mass Flows:

SD-CCVC Isochoric/Adiabatic Cycle:

1 mole H2 equals 2.0 grams. 0.015 moles equals 0.03 grams. Assuming thelow heat of combustion, 1 kg of H2 has a combustion value of ˜120,000 kJ(33.33 kW-h), or 120 kJ/gram. For the SD-CCVC isochoric/adiabatic cycle,which requires 3,726 kJ/hr, that would require 31.05 grams of H2/hrbeing converted to H2O. O has 8× the mass of H2. Therefore, 248.4 gramsof O2/hr would be required.

SD-CCVC Isobaric/Adiabatic Cycle:

For the SD-CCVC isobaric/adiabatic cycle, 3,284 kJ/hr, 27.4 grams ofH2/hr and 219 grams of O2/hr would be required.

SD-CCVC Isothermal Cycle:

For the SD-CCVC isothermal cycle, 4,498 kJ/hr, 37.5 grams of H2/hr and300 grams of O2/hr would be required.

Per hour, (38.4/2034=) 1.89% of the total cycled H2 working fluid willbe combusted.

The BB Closed Loop Process.

The BB closed loop process essentially utilizes a thermochemicalC6H12<=>C6H6+3H2 cycle similar to that disclosed in U.S. Pat. No.3,225,538, but configures it differently, seeing the H2 generated bydissociation of a cyclical hydrocarbon such as C6H12 as “linked” with asecond thermochemical and/or electrochemical cycle that associates anddissociates H2O. FIG. 22 below is based on FIG. 1 in U.S. Pat. No.3,225,538, with temperatures in degrees Kelvin and pressure inatmospheres measured logarithmically to the base 10 (an insert graphsthe base 10 into the righthand bottom of the figure). Per FIG. 22 , acyclical cyclohexane<=>benzene+hydrogen catalytic process at a giventemperature and pressure will be either endothermic or exothermic. Forexample, at a pressure of 4 atm, the temperature for 90% endothermicconversion equals ˜900 K (1,620 R), while the temperature for 90%exothermic conversion equals ˜720 K (1,296 R).

Many forms of thermal energy can be used to dissociate the C6H12 intoC6H6+3H2. However, a particularly interesting approach has been proposedwhereby the required thermal energy can be generated by oxidizing aquantity of the H2 thus released.

1 kg of H2 has a mass of 2.20 Lb and thus a low heat value of 120,000kJ/kg (33.3 kWMh, 44.7 HP/hr). Since ˜1,062 kJ are absorbedthermochemically in 1 Lb (0.4536 kg) of C6H12, or 2,341 kJ/kg, thatmeans the combustion of 1 kg of H2 can theoretically convert 51.3 kg ofC6H12 into C6H6+3H2. Since C6H6 has a mass of 78.11 g/mol and C6H12 hasa mass of 84.16 g/mol, the conversion yields 6.05 g of H2 per mol ofC6H12. 51.3 kg equals 609 mols of C6H12. Therefore, the mass of H2released equals ˜3.7 kg.

That is, 27% of the 3.7 kg H2 released is burned to produce 73% or 2.68kg of H2. Looking at the information under the heading “Typical SD-CCVCcycle H2 and O2 mass flows” above:

The SD-CCVC isochoric/adiabatic cycle thermal requirement equals 31.05grams of H2/hr, or 1.12% of the H2 freed from 51.3 kg of C6H12.Therefore, total C6H12 required to fuel the cycle equals 0.59 kg perhour.

The SD-CCVC isobaric/adiabatic cycle thermal requirement equals 27.4grams of H2/hr, or 1.0% of the H2 freed from 51.3 kg of C6H12.Therefore, total C6H12 required to fuel the cycle equals 0.524 kg perhour.

The SD-CCVC isothermal cycle thermal requirement equals 37.5 grams ofH2/hr, or 1.40% of the H2 freed from 51.3 kg of C6H12. Therefore, totalC6H12 required to fuel the cycle equals 0.72 kg per hour.

The SD-CCVC displacement expansion cycle thermal requirement willresemble the SD-CCVC isochoric/adiabatic cycle thermal requirement. Itwill be assumed that total C6H12 required to fuel the cycle willlikewise approximate 0.60 kg per hour.

In a BB closed loop process, H2 and O2 will be generated from H2O, as byelectrolysis. (Note: H2 and O2 can also be generated from H2O by thermalcracking.) The H2 is then used to convert C6H6 into C6H12, and alsogenerate useful heat, as discussed below. Converting H2 into C6H12permits storage of the H2 as a liquid at ambient pressure andtemperature. The O2 may also be stored, either as a pressurized gas, asliquified oxygen (LOX), or within some form of thermochemical carrier.(One potential thermochemical carrier is hydrogen peroxide (H₂O₂). Notethat H₂O₂ can also be stored as a liquid at ambient pressure andtemperature.)

The C6H12 and possibly stored O2 are then transported to the powerproduction site. At the site, the C6H12 will be run through anendothermic catalytic reactor, absorbing thermal energy and producingC6H6 and H2.

As noted above, one possible thermal energy source for converting C6H12into C6H6+H2 is combustion of about a quarter of the H2 thus converted.U.S. patent Ser. No. 18/095,463 (applied for) illustrates severalprocesses that can be made to produce net W-out from a C6H12>C6H6+3H2reaction. In fact, such work-generating cycles were envisioned by theoriginal B/E Cycle inventors, Reginald Bland and Frederick Ewing, andare disclosed in U.S. Pat. Nos. 3,225,538, 3,067,594, and, posthumously,U.S. Pat. No. 3,871,179.

It is anticipated that an SD-CCVC cycle engine, specifically designedfor the main purpose of efficiently generate H2 from the catalyticdissociation of C6H12, and for a secondary purpose of efficientlygenerating W-out, may be powered by the combustion of 27% of the H2 gasthus produced. Such an SD-CCVC variant would be externally heating, bycombustion of H2, an endothermic catalytic reactor. C6H12 would bedissociated into C6H6+3H2, creating half of a classic “Bland/EwingThermochemical Cycle” that can take full advantage of the molecular 4:1expansion therein, first at constant pressure and temperature expansion,and then at adiabatic expansion. The exhaust from what is shown as the1st stage compressor in FIG. 11 through FIG. 17 would principally beC6H6 and H2 gas (and some remnant C6H12). As will be shown below, hotC6H6+H2 exhaust can be directed into a C6H12 “vaporizer” heat exchanger.A new “charge” of C6H12 would be pumped into the other side of the heatexchanger at a lower pressure than the exhausting C6H6+H2 mix. Higherpressure C6H6 would condense at a higher temperature than that requiredto vaporize the liquid C6H12, thus utilizing the heat of condensing C6H6for most of the required thermal input for the C6H12 vaporizationprocess. The vaporized C6H12 would then enter a multistage and carefullyintercooled compressor (ensuring that the vaporous content avoidedcondensation), returning the C6H12 vapor to the desired pressure forreceipt by the lower displacer while approaching an isothermalcompression. The lower displacer would then cyclically displace thepressurized C6H12 vapor through the regenerator at constant volume.Meanwhile, the condensed C6H6 (and C6H12 remnant) is easily separatedfrom the H2, some of the separated H2 being then potentially used asfuel to “power” the engine, and the remainder of the separated H2 thenbeing available for other uses.

(Note: A possible side benefit of using such a working fluid would bethe possibility of using a small amount of liquid C6H12, or possiblyC6H6, to help lubricate the engine. That in turn may permit higheroperating temperatures. Any lubricant “leakage” would simply be carriedthrough the engine and recaptured in the exhaust.)

One simple system for storing and delivering C6H12 is shown as FIG. 23in this application. Note that W-in to the hydraulic pump is the totalW-in required to feed the system, and that a large degree of therequired hydraulic pump W-in is returned as W-out by the hydraulic motorshown receiving the dissociated C6H6 in pressurized liquid form. Whilethe system shown in FIG. 23 appears to be very wasteful of energy, notethat the output is not work, but high pressure H2 gas. In effect, the“work out” is H2 that is in effect “pressurized” by the endothermicconversion process, with a given pressure being determined by thetemperature at which the endothermic conversion is made to take place.With H2 combustion, very high peak temperatures are easily achieved,thus very high pressures can be achieved as well. In this sense, thecatalytic dissociation of C6H12 into high pressure C6H6 and H2 can beseen as a potential means of thermochemically compressing H2 gas by ameans that bypasses the need for mechanical compression. This isparticularly useful in potentially high pressure heat engines such asthe SD-CCVC designs proposed herein. Here are the steps:

-   -   a. Liquid reactant (C6H12) at ambient temperature and pressure        is pumped out of storage, for example in a cylinder possessing a        double-acting piston, on one side of which is stored the liquid        reactant and on the other side of which is stored the liquid        product (C6H6 and remnant C6H12).    -   b. The liquid reactant is pumped to 100 atm and exhausted into a        recuperator (or regenerator such as a valved regenerator.    -   c. The recuperator which will preheat, vaporize, and superheat        the reactant to approximately the temperature of the endothermic        reactor. To accomplish this preheat, it will transfer heat to        the reactant from the product exhausting from the endothermic        reactor.    -   d. The endothermic reactor will convert the reactant (C6H12        superheated vapor) into the product (superheated C6H6+H2 and        remnant C6H12) with heat it receives from a high temperature        heat source. That high temperature heat source can be a portion        of the H2 produced and/or excess heat generated by an operation        such as the combustion of H2.    -   e. The product is flowed to the other side of the recuperator or        regenerator.    -   f. The product is cooled under pressure sufficiently to separate        the H2 (and other gases, if any) from the C6H6 and remnant C6H12        (and other liquids, if any).    -   g. The liquid products are flowed through a hydraulic expander,        reducing them to ambient pressure.    -   h. The liquid products are sent to the other side of the        double-acting piston, where they are stored. The stored products        will eventually be exchanged for a fresh charge of reactant.    -   i. The high pressure gases are flowed to their destination, for        example to a reheater (either constant pressure or constant        volume) and then to an H2-burning heat engine.

(It is also possible to envision such a system for convertingpressurized H2O2 into H2O plus highly pressurized O2. However, in H2O2,the dissociation reaction is highly exothermic. Under the circumstances,it might be best to simply inject pressurized H2O2 directly into the H2stream, possibly without any preheating.)

Following endothermic conversion, the C6H6 and H2 product will be cooledto the point where the C6H6 is liquified and separated from the H2 gas.The H2, and possibly previously stored O2. are then converted back intoH2O, either in a fuel cell or by combustion, essentially releasing theenergy stored when they were generated in the first place.

FIGS. 24A, 24B, and 24C illustrate how a fluid, for example liquid C6H6[1], can be captured in a tank [2] that is holding, for example, liquidC6H12 [3], as the original C6H12 is “emptied” and converted into C6H6,showing top, middle, and bottom positions of the “charging piston”.

Illustrated in FIGS. 24A, 24B, and 24C is a simple “charging” piston [4]and cylinder [2] arrangement. As will be shown, this chargingarrangement permits charging or discharging of fluid [1 and 3] movinginto and out of the displacer cylinder. In this instance, a connectingrod [5] is shown running through the charging cylinder, connecting tothe double-acting charging piston internal to the charging cylinder(charging cylinder connecting rod seals are not shown). Note: The roddoes not have to be double-sided, but doing so illustrates that anypressure acting on one side of the piston is perfectly equal to thepressure acting on the other side (barring friction).) The double actingcharging piston is shown with a sealing ring [6] the sealing ringarrangement around its circumference, which can allow a differentiationof the fluids on either side of the piston. The charging piston is usedto create the direction of constant pressure fluid flow, being typicallymoved by some small force in one direction or the other. It thusrepresents a simple directional pump.

From U.S. patent application Ser. No. 18/197,902, section“Specification—Miscellaneous Descriptions and Operations” entitled “TheBenzene Battery Cycle”: “It is obvious that the BB cycle energy storageand delivery process has a potential usefulness beyond the lunarsurface. In fact, it can easily be shown to represent a meaningfulalternative to the present hydrocarbon-combustion processes thatcurrently underpin much of the human race's energy generation anddistribution network.” The application then goes on to describe thefollowing means: “The service station fills a transport's tank withC6H12 while emptying the same tank of C6H6 (a partition keeping the twoliquids separate from one another).

It is anticipated that the generated C6H6 (plus any remnant C6H12) [1]would then be returned to the tank [4] holding the C6H12 [2], with thetwo liquids being kept separate, as by the use of a piston. For acomplete system, a vehicle receives a fresh charge of C6H12 at a“service station” in exchange for the liquid C6H6 and remnant C6H12. Thereceived liquid is then shipped back to a H2 production plant forconversion back into C6H12.

C6H6 as a Means for Capturing and Transporting Potential Thermal Energy.

In supplying the thermal energy to release the H2 gas from the C6H12,thereby converting the C6H12 to C6H6, note that the total amount ofdissociation thermal energy required is “captured” chemically within theC6H6 (assuming a perfect thermal recuperation, see FIG. 23 ). Thatthermal energy will be “released” exothermically when the C6H6 isconverted back into C6H12 by the absorption of a new supply of H2. Notethat the temperature of the thermal release can be increased ordecreased by simply adjusting the pressure of the C6H6+3H2 mix prior tothe mix entering the exothermic reactor, as shown in FIG. 22 . In theBenzene Battery closed loop system, upon being returned to the thermalsource power plant (such as a solar thermal power plant), the thermalenergy generated by the recombination of H2 and C6H6 can be madeavailable to the thermal power plant, thereby increasing its overallefficiency.

A simple heat generating system is shown in U.S. patent Ser. No.18/095,463, FIG. 14 , shown as FIG. 25 in this application. This systemmakes use of a technique termed an Exothermic Reactor Exhaust Compressor(EREC), disclosed in U.S. patent Ser. No. 18/095,463. In essence, theEREC, by raising the relative pressure of the C6H12 and remnant C6H6exiting the exothermic reactor, permits thermal energy of condensationfrom that C6H12 and remnant C6H6 to supply much if not all of therequired energy of vaporization for the inflowing liquid C6H6. Theresult is the generation of a net thermal energy output of 90 kJ from aconversion of 0.038 kg of C6H6+H2 to C6H12 which yielded 98.3 kJ ofthermal energy. The cost, which is essentially the W-in required tooperate the EREC, is a theoretical power input of about 1.2 kJ.

U.S. patent Ser. No. 18/095,463 also proposes several alternativeapproaches termed “Exothermic production cycles. There are three usecases for this thermochemically produced heat; cogeneration or combinedheat and power (CHP), combined cycle (CC), and hybrid CHP/CC. The usecases differ primarily in the relative percentage of power output versusthermal output.

It is also possible to use the released exothermic thermal energy topower, or aid in powering, an SD-CCVC cycle engine. In FIG. 22 , it isshown that an exothermic reaction at only ˜50 psi (345 kPa) willgenerate well over the 555 K (1,000 R) temperature regime explored inthe above calculations. Thus, the C6H6+H2 reaction itself can easilypower an SD-CCVC cycle engine. A very low temperature SD-CCVCdisplacement expansion engine, for example, has been shown herein totheoretically generate power at a thermal efficiency in the range of16%, and much higher peak temperatures and hence efficiencies appear tobe possible.

Finally, another alternative that uses an SD-CCVC cycle involvesreplacing the O2 injected in the engine designs described herein withvaporized C6H6, while leaving the pressurized working fluid as H2.Generally speaking, an exothermic catalytic reactor would be added justbefore the SD-CCVC expander, replacing the manifold labeled “injectorsite” shown in FIG. 16 and FIG. 17 . Vaporized and pressurized C6H6would be cyclically injected into a primarily H2 working fluid atapproximately the point where working fluid is passing out of the upperdisplacer cylinder. The working fluid mix would then be passed throughthe catalytic reactor, endothermically converting the C6H6 to C6H12,thus supplying the thermal energy to drive the engine by internalcatalytic heating. The C6H12 would then be liquified and removed fromthe working fluid at approximately the point where H2O is shown removedin FIG. 15 and FIG. 17 .

The SD-CCVC cycle proposed for use with this alternative is the SD-CCVCisobaric/adiabatic expansion cycle. A constant pressure and temperatureheat addition needs to be maintained during the injection of thevaporized and pressurized C6H6 as it passes into the exothermiccatalytic reactor, although there is a possibility of a small amount ofisochoric exothermic chemical heat addition slightly in advance of theexpansion. Such a cycle can be termed a benzene-fueled SD-CCVCisobaric/adiabatic cycle.

As in SD-CCVC cycles described above, the H2 gas will be “densified”using multi-staged and intercooled compressors to approach isothermal asclosely as practicable. Waste exhaust heat will be added via the STREPprocess. And a highly pre-pressurized H2 working fluid is also possible,increasing power density. Pre-pressurizing the H2 working fluid willalso increase the temperature of the catalytic exothermic recombinationof C6H6 vapor and H2.

Ideally, pressurized C6H6 will be vaporized using the EREC process asshown in FIG. 25 and described in U.S. patent Ser. No. 18/095,463. Notethat care will need to be taken to exhaust the working fluid mix fromthe 1st stage compressor before the vaporous C6H12 begins to condenseout. The working fluid mix would then exhaust into a C6H6 “vaporizer”,utilizing the heat of condensing C6H12 for most of the required thermalinput for the vaporization process. This process is described in detailin U.S. patent Ser. No. 18/095,463. That would permit the C6H6 to bepumped into the system at high pressure, greatly increasing overall netW-out and thermal efficiency.

In designing a prototype SD CCVC engine that best utilizes the existingCCVC prototype, several new applications and improvements have beendiscovered.

Note: To develop a first order analysis for various proposed cycles, anideal gas law calculator and the CGL calculator will be used. To preparea visual estimate of various cycles, these calculations will be used inconjunction with FIG. 18 , taken from FIG. 70 , “Marks' MechanicalEngineers' Handbook”, 6th edition, 9-148, “Internal-combustion engines”.FIG. 27 shows lines traced onto FIG. 18 and then moved from the tracingfor clarity purposes, representing various states of working fluid forvarious cycles. FIG. 27 also includes letters representing variouspoints for the various states. See “1st Cycle analysis”, paragraph 10,below. The solid vertical lines in FIG. 27 represent zero heat transfer(isentropic/adiabatic) changes, the solid non-vertical curves representconstant volume (isochoric) changes, and the dotted non-vertical linesrepresent constant pressure (isobaric) changes. In FIG. 18 and FIG. 27 ,temperatures are in degrees Rankine, pressures are in pounds per squareinch (psi), energy and enthalpy are in Btu's, and volumes are in cubicfeet. To aid in use of the ideal gas law and CGL calculators,temperatures will also be shown in degrees Kelvin, pressures inkiloPascals, energy in kiloJoules, and volumes in liters. Heat and powerwill be shown in kiloJoules and Btu's.

Calculations herein concerning endothermic and exothermic dissociationand re-association of a mol of cyclohexane (C6H12) into a mol of benzene(C6H6) and three moles of hydrogen (H2) will utilize FIG. 1 from U.S.Pat. No. 3,225,538, shown as FIG. 22 herein. As shown in FIG. 22 , inthe presence of a proper catalyst C6H12 will move towards C6H6+3H2 iftemperature remains constant and pressure is increased and in theopposite direction if pressure decreases. Likewise, if pressure isconstant, C6H12 will move towards C6H6+3H2 if temperature is decreasedand in the opposite direction if it is increased. FIG. 22 indicates thepercentage of conversion in a given direction, with the line marked,“99” representing a 99% C6H12 content and the line marked 0.01representing a 99% C6H6+3H2 content. Finally, note that this particularreaction is highly stable and thus highly cyclical. However, it has beenestablished that some degradation will occur over time, with conversioninto other forms of hydrocarbon.

Finally, note that all proposed cycles and mechanisms are being shownsolely to describe the inherent nature of the broader inventions beingherein disclosed. Thus, in FIGS. 11 through 17 , the expander cylinder,the synchronizer, the compressor or compressors, and the displacers areshown as separate modules that interact. Alternatively, as will be shownherein, a single crankshaft throw may be used for all modules, with thevarious drive rods of the various modules physically connected. Also,shown embodiments of the proposed cycles, such as starting temperatures,pressures, volumes, proposed working fluids, etcetera, do not innatelydescribe ideal embodiments but are to be considered first orderestimates. Much actual reduction to practice and experimentation will berequired to determine that.

Additional Improvements:

-   -   (1) In some versions, as will be shown, the isochoric exhaust        displacer or “synchronizer cylinder” can be easily converted        into a secondary expander cylinder, and thus will have a larger        volume than the primary expander cylinder. Note that, to permit        true isochoric exhaust displacement, a third cylinder, that is,        the cylinder that receives the exhaust from such a secondary        expander cylinder, will have the same volume as the secondary        expander cylinder.    -   (2) A secondary addition of source heat following the primary        expansion of the working fluid is also possible, as will be        discussed below. Such additional heat may be added prior to        exhausting into a secondary expander cylinder. Note that        additional source heat can be added at constant volume,        pressure, or temperature, or at some combination thereof. Also        note that such a secondary heat addition is well known to those        skilled in the art to be thermodynamically beneficial.    -   (3) It has been found to be useful to have two or more isochoric        thermal input displacement means performed in series. For        example, the first displacement can utilize lower temperature        exhaust waste heat. A second displacement can then utilize        higher temperature H-in, for example from a thermochemical        catalytic reaction, such as the exothermic catalytic reformation        of benzene (C6H6) plus hydrogen (H2) into cyclohexane (C6H12).        In one such means, the cylinder presently shown as the 2nd stage        compressor cylinder in FIGS. 11 through 17 can be used instead        as a new lower displacer cylinder, the lower displacer cylinder        shown in FIGS. 11 through 17 will become the middle displacer,        and the upper displacer shown in FIGS. 11 through 17 will remain        the upper displacer. Between the lower and middle displacer        cylinder will be the lower temperature regenerator, and between        the middle and upper displacer cylinder will be the higher        temperature regenerator. As will be shown, two different        approaches have been considered for thermally charging the        higher temperature regenerator with heat produced by an        exothermic reactor. The first approach would heat a thermal        transfer fluid (such as He or H2) via a heat exchanger used to        cool the exothermic reactor, said transfer fluid then being        intermittently flowed through, and thus “charging”, the        regenerator. The second approach, perhaps more useful for        generating useful work, would add excess cooler thermal transfer        fluid at the same pressure to the endothermic fluid as it is        being created by said exothermic reaction, essentially cooling        the reactor internally. This may require multiple injection        points along the reactor body to avoid over-cooling.    -   (4) Rather than performing “displacement expansion” or        non-adiabatic expansion by exhausting from the final displacer        through a heater and directly into the expander cylinder as        proposed in FIGS. 11 through 17 , the working fluid exhausting        from the final displacer cylinder can isobarically exhaust. This        permits “decoupling” of the displacer system from the expander        system so that the displacer system and the expander system        don't have to be phase-locked to one another. It also permits        isobaric source heat addition, allowing a large distance to        exist between the final isochoric displacer and the expander.        That in turn would permit, for example, a thermal solar energy        system of large volume, such as a trough-type solar        concentrator, to be used as the thermal source for the heat        engine. The primary expander can then utilize, for example, a        combination isobaric/adiabatic expander, akin to a diesel cycle,        whereby:    -   A. During the first portion of expansion, a “charge” of working        fluid exhausted from the upper displacer at constant pressure        will be taken into the primary expander at constant pressure via        the existing intake/transfer valve.    -   B. During the second portion of expansion, the charge of        constant pressure working fluid thus captured will then be        expanded adiabatically. For example, following a degree of        isobaric expansion, the intake/transfer valve can be        instantaneously closed, as by the action of a solenoid, thus        instantaneously converting the expansion into an adiabatic one.    -   (5) Isobaric heat addition is useful, but heat may also be added        isochorically, isothermally, or some combination of the three.        Note that, in the graphs indicating working fluid states shown        in FIG. 27 , multiple possible configurations are shown,        including isobaric and isochoric configurations. Isothermal        source H-in is also possible.    -   (6) A “valved cell” can be the means for attaching pre-heated        and/or pre-pressurized fluids into the expander cylinder, via        the concepts disclosed in U.S. Pat. Nos. 4,817,388 and        5,179,839.    -   (7) STREP heat exchangers can be seen as applicable to what can        be termed the Bland/Ewing Composite Cycle (B/E-CC) process. As        stated in Pending U.S. patent Ser. No. 18/095,463: “The        underlying improvement to the foundational B/E Cycle invention        disclosed in U.S. Pat. Nos. 3,225,538, 3,067,594, and 3,871,179        takes the form of apparatuses for combining various        independently operating or semi-independently operating        endothermic and exothermic half-cycles which coupled together        form a complete B/E Cycle.    -   (8) In U.S. patent application Ser. No. 18/095,463, use of an        Exothermic Reactor Exhaust Compressor (EREC) is proposed to        assist in the vaporization of C6H6 by a counter flowing exchange        of heat with condensing higher pressure C6H12. It is herein        proposed that an Endothermic Reactor Exhaust Compressor (which        will be termed an ENREC) be used to permit the condensation of        higher pressure C6H6 to supply some or all of the thermal energy        required to vaporize a similar mol count of C6H12.    -   (9) FIGS. 11 through 17 proposed increasing the efficiency of a        isochoric counterflow recuperator, such as was used in the        original Closed Cycle Valved Cell (CCVC) prototype as described        therein, by using (1) an expander exhaust receiver cylinder of        equal volume to said expander, termed an SD cylinder, (2)        valving and connecting manifold means that intermittently and        cyclically connected said SD cylinder to (3) a regenerator,        and (4) valving and connecting manifold means that        intermittently and cyclically connected said regenerator to an        equal volume second displacer cylinder. Note that such a means        meets the general definition of a STREP. Various additional        modified STREP designs are herein proposed.

For example, a modified STREP can be used as means of exchanging heatbetween a low pressure high volume “receiver” mechanism and a highpressure low volume “displacer” mechanism, as shown in FIGS. 1 through 4. Note that the higher volume flow of the “heating” fluid would be dueto the difference in density required for a given amount of heat to betransferred. It can be desirable for a high temperature but low pressureexhaust fluid (for example, exhaust from a heat engine) to be passed atconstant pressure through a counterflow regenerator-type heat exchanger,thermally “charging” the regenerator. The regenerator can then be raisedin pressure, for example by a semi-adiabatic compression of remnantfluid in the “receiver” mechanism such as would be caused by an earlyclosure of the regenerator's exhaust valve. A separate stream ofcounter-flowing fluid at the higher pressure but at low temperature canthen enter the regenerator through an intake valve and flow isobaricallythrough the regenerator via the displacer mechanism, thus isobaricallyabsorb the thermal energy deposited by the earlier low pressure flow.Finally, the high pressure in the regenerator can be reduced to that ofthe low pressure stream and once more be used to “charge” theregenerator, as by re-expansion of remnant high pressure fluid in theregenerator back into the low pressure displacer at the beginning of itsintake stroke, followed by low pressure fluid intake into said lowpressure displacer.

In addition, the high pressure, high temperature fluid thus generatedcan then be used to isobarically “charge” an isochoric regenerator up tothe peak temperature of the low pressure fluid. That is, a modifiedSTREP may be used to thermally charge a “valved regenerator” with anisobaric fluid, then “switch” the regenerator via valving in order toisochorically remove some or all of the thermal charge thus depositedisobarically. Note that this is ideal for capturing high temperaturethermal energy from the low pressure exhaust of a heat engine. Forexample, a high pressure, high temperature fluid can be passed atconstant pressure through a counterflow regenerator-type heat exchanger,whereby the exhausting fluid's thermal energy can thus thermally“charge” the regenerator. A separate stream of counter-flowing fluid atthe same pressure but at low temperature can then enter the regeneratorthrough an intake valve. If the flow were made to be isochoric, thecounter-flowing fluid would absorb the thermal energy deposited by theearlier flow at constant volume, as by displacement from an initialcylinder, through the regenerator, and into an SD cylinder, thustheoretically both raising the temperature of the counter flowing fluidto the peak temperature of the thermal charging fluid, and thermallyrather than compressively raising the pressure of the counter-flowingfluid. Following the thermal charging of the higher pressurecounter-flowing fluid, the SD cylinder exhaust valve would close and theremnant gas in the regenerator would be expanded back into the displacercylinder until it dropped to the pressure of the thermal charging fluid,at which point the regenerator intake valve would close, the regeneratorexhaust valve would open, and the cycle would begin again.

1st Cycle Analysis

FIGS. 29 a through 30 c show different cutaway views of a CAD solidmodel of one possible approach to converting the existing CCVC prototypeengine into a prototype engine with the above characteristics. Severalassumptions are made:

-   -   (1) Initially, the existing prototype will be used as much as        possible, which “locks in” certain volumes and temperatures.        These are defined above and associated with FIGS. 11 through 18        .    -   (2) Ideally, the proposed engine will use multiple intercooled        compressions. A “sink” or low cooled temperature of        approximately the boiling point of water is assumed.    -   (3) The proposed engine will assume use of a single expander,        although a second expansion is possible, as proposed above.    -   (4) In the proposed prototype engine, there will be two        isochoric displacements through two different heaters via three        displacer cylinders. The upper displacer piston, which will        operate in a higher peak temperature environment, will be formed        with an upper piston extension which will permit the upper        piston seal to mount a teflon/spring seal in the upper cylinder        wall. Mounting the seal in the cylinder wall will allow cooling        of the seal via the cylinder wall area that is in physical        contact with the seal, permitting a potentially higher working        fluid temperature.    -   (5) In the proposed prototype engine, there will be an isobaric        displacement through a third heater into an expander valved cell        working fluid injector. There will then be an adiabatic        displacement via the valved cell expander injector into the        primary expander. That will permit testing of additional cycles,        as will be shown.    -   (6) Once in operation, the proposed prototype is expected to be        capable of testing at much higher base pressures, much as a        classic stirling engine may be operated at very high base        pressures. Note that this “pre-pressurization” will not        measurably impact the various temperatures within such a        pre-pressurized engine, just as in a classic stirling engine.        What will be impacted is power density per cycle, which is an        extremely important element in achieving workable power outputs        and thus potentially approaching theoretical potential. That is        because, for any set of given inefficiencies from friction,        pumping, or thermal losses, a higher power density decreases        these losses relative to the actual net power produced.

In the proposed prototype engine, it will be initially assumed that thevolume of each of the three displacer cylinders will be comprised of a2.5″ diameter cylinder containing a 2″ diameter connecting tube over adisplacer piston stroke of 2.75″, or an effective piston area of 1.767sq in and thus a cylinder volume of 4.86 cu in (0.0796 L).

As previously stated, FIG. 27 includes letters representing variouspoints for the various working fluid states. These will be referenced asthis analysis proceeds.

It is estimated that, at 671 R (373 K, 100 deg C., 212 F) and a pressureof 1 atm, the volume is increased to 497 L/min (17.6 cu ft/min) (PointA). Following stage one compression (Point B) and intercooling (PointC), per the CGL calculator, the 497 L/min charge of working fluid willbe assumed to be compressed to about 25 psi (172.4 kPa) at 671 R (373K), decreasing volume to 353.7 L/min (12.5 cu ft/min). Following stage 2compression (Point D) and inter cooling (Point E), the charge isreturned to 373 K (671 R) at a pressure of about 74 psi (510 kPa),dropping the volume per the ideal gas calculator to 103.8 L/min (3.667cu ft/min, 6,337 cu in/min).

Beginning at BDC, the lower displacer cylinder will receive a constantpressure charge of 74 psi (510 kPa) working fluid at 373 K (671 R). Thatwill equal 1/1,304th of the 17 cu ft/minute flow, indicating that theengine would need to have a rotational speed of 1,304 rpm to passthrough 17 cu ft/minute. Per the CGL calculator, at a final temperatureof 373 K the final flow volume per minute into the lower displacerequals 3.67 cu ft (103.8 L), or 4.86 cu in/cycle (0.0796 L/cycle) @1,304 rpm), internal energy change equals 18.5 kJ, W-in (of exhaust)equals 8.1 kJ (7.6 Btu), and rejected heat equals 27.9 kJ. Per the idealgas law, the H2 flow rate equals 17.07 moles/min or 0.0131 moles/cycle.Since 16.24 moles pass through every minute, the total charge inductedinto the lower displacer cylinder would equal 0.0124 moles. At STP, H2(gas) has a mass of 2.02 g/mol. 0.0124 moles of H2 thus equals 0.026 gor 26 mg.

At TDC, following intake of 0.0131 moles of working fluid at 373 K and172 kPa into the lower displacer cylinder, the working fluid will beexhausted completely into the #1 regenerator and the middle displacercylinder. Since the two displacer cylinders have the same displacementand the same stroke, the displacement occurs isochorically. Sincethermal energy is added, the temperature of the working fluid followingisochoric displacement via the following move to BDC will ideally reachthe temperature of the exhaust fluid charging the #1 regenerator, or 840R (466 K). Per the CGL calculator, the pressure will increase to about648 kPa (94 psi) (Point F).

Note that the thermal energy source may be otherwise-waste heat from theengine exhaust.

During the following move to TDC, the middle displacer will then exhaustcompletely into the #2 regenerator and the upper displacer cylinder.Since the two displacer cylinders have the same displacement and thesame stroke, the displacement occurs isochorically. Since thermal energyis added, the temperature of the working fluid following isochoricdisplacement via the following move to BDC will reach the approximatetemperature of the exhaust fluid charging the #1 regenerator, or about984 R (547 K). Per the CGL calculator, the pressure will increase toabout 758 kPa (110 psi) (Point G).

Note that the thermal energy source may be thermochemical heat releasedby the conversion of C6H6 plus H2 into C6H12.

In the following model, the temperature of the proposed prototype willnot exceed 1,140 R (633 K, 360 deg C., 680 deg F.). That temperaturewill be limited to the primary expander and the primary expanderconstant pressure injector, as will be shown. To accomplish this, sourceheat will be added at constant pressure. The heat exchanger exterior tothe upper displacer will raise the working fluid temperature to 1,140 R(633 K).

The isobaric H-in will occur as the working fluid stream exhausted atconstant pressure from the upper displacer is subsequently heated to ahigher temperature. Temperature during constant pressure upper displacerexhaust will occur at about 547 K (984 R). Consequently, the upperdisplacer exhaust will be through a high temperature exhaust checkvalve. A non-lubricated externally cooled teflon bearing exhaust checkvalve guide should be possible if it is sufficiently physically removedfrom direct contact with the working fluid. In addition, to allow thishigher temperature without using a lubricant, the upper displacer pistonwill be lengthened, such that a teflon/ss spring seal can be mounted inthe cylinder wall. With external cooling of the cylinder wall inimmediate proximity to the teflon seal and with the standoff distance ofthe lengthened piston, the upper displacer piston will likely notrequire lubrication.

The initial receiver for the 633 K, 110 psi working fluid will be a“valved cell”. The physical nature of the cell will be that of a pistonand cylinder, with the piston sharing the 2.75″ stroke of the overallengine. It will be positioned on top of the expander cylinder and itscharge of working fluid will be connected to the expander cylinder via a“transfer valve” which will operate similarly to that on the existingCCVC prototype, as will be shown. The valved cell will also possess amechanically operated high temperature inlet valve. Non-lubricatedexternally cooled teflon bearings for the intake and exhaust check valveguides should be possible if the valve guide seats are sufficientlyphysically removed from direct contact with the working fluid. Betweenthe upper displacer exhaust check valve and the valved cell intakevalve, an isobaric high temperature heat source heat exchanger will besituated. It may be partially or completely composed of the existingCCVC prototype heater and/or partially or completely composed a solarconcentrator such as the original Bland/Ewing Cycle parabolic troughsolar concentrator. As noted above, volume equals 3.80 cu ft at 1,140 R,closely matching the relevant graphed line in FIG. 27 .

In operation, the valved cell would act as a constant pressure expanderas it is charged via its intake valve, as the valved cell piston ismoving from BDC to TDC. The transfer valve into the expander would thenbe opened at or around TDC, and the working fluid would exhaust in anadiabatic displacement expansion into the expander piston, resulting inrapidly falling pressure and temperature. Towards the end of theexpansion stroke, the transfer valve would close, allowing a smallamount of gas to be trapped and pressurized up to 110 psi. On the moveback to TDC, the valved cell intake valve would then be opened, allowingthe valved cell to be charged once again with H2 working fluid at 110psi and 1,140 R (633 K) (Point H).

To allow this higher temperature without using a lubricant, both theexpander and the valved cell pistons will be lengthened in the samemanner the upper displacer piston is lengthened, such that anon-lubricated teflon/ss spring seal can be mounted in the cylinderwalls of each. Note that, with external cooling of the two cylinderwalls in immediate proximity to the teflon seals and with the standoffdistance of the lengthened pistons, the valved cell displacer and theexpander will likely not require lubrication at these temperatures.

To compensate for the potentially significant added mass of the muchlonger pistons, the crankcases of the upper displacer, valved cell, andexpander cylinders will be pressurized to a higher pressure than thepeak pressure developed within the three cylinders, and be held atconstant pressure. Consequently, the piston walls can be made quitethin, and piston mass can be held down appreciably. Note that it ispossible to accomplish some cooling of the internal portions of thethree pistons by this compressed crankcase gas, thus permitting a higherpeak temperature for the non-lubricated engine.

In the primary expander, which has an expanded volume of 13.5 cuin/cycle (0.221 L/cycle), or 288.5 L/min (10.19 cu ft/min), an adiabaticexpansion process will be assumed. A charge per cycle, equal in mass tothe charge exhausted from the upper displacer, will be injected into theexpander via displacement. Dead space at the completion of the expanderexhaust will be considered nearly equal to zero. That equates to anexpansion ratio of 13.5/4.86 or 2.777 to 1 (Point I).

The expander charge would then essentially completely exhausted atconstant pressure and temperature into a SD cylinder, which has a closevolumetric match to the expansion cylinder. Since pressure andtemperature are maintained, no work either in or out is done.

Following synchronizer piston intake, the charge will be essentiallycompletely exhausted into a displacer/1st stage compressor cylinder withmatching piston diameter and stroke. That creates an essentiallyisochoric displacement process that can be timed to pass through andthus charge the regenerator between the lower and middle “compression”displacer. Note that this can be thought of as a kind of constant volumedisplacement “decompression”, since the working fluid pressure is lowerfollowing displacement due to the removal of heat by the regenerator.The result is the charging of the regenerator with waste heat from theexpander and the simultaneous cooling of the isochorically-displacedcharge exhausted by the synchronizer, dropping its temperature toapproximately 671 R (373 K) at a pressure of about 25 psi (172.4 kPa).Volume remains 10.2 cu ft (Point A).

W-in, W-out, H-in, and theoretical thermal efficiency can now beestimated.

Exhaust from the expander from BDC now proceeds at constant pressure,temperature, and volume by means of the synchronizer cylinder. Since allstates remain unchanged, the only negatives are friction, pumpinglosses, and thermal leakage, which are assumed to equal zero. By TDC,the expander has displaced all but a small amount of clearance gas intothe manifold and from thence into the synchronizer cylinder. Closure ofthe expander exhaust valve slightly early allows the expander clearancegas to be pressurized to equal that of the gas that, in the previousstroke, charged the valved cell injector, opening the transfer valve atapproximately TDC.

The synchronizer piston now exhausts to BDC through the valvedregenerator and into the 1st stage compressor at constant volume,charging the regenerator with heat and isochorically cooling the gasdisplaced from the synchronizer and into the 1st stage compressor. Perthe CGL calculator, volume equals 9.71 cu ft (275 L) and stays constant,pressure drops from 36.5 psi (252 kPa) to 29.2 psi) (201 kPa), andtemperature drops from 466.25 K (840 R) to 671 R (373 K),releasing/storing 34.1 kJ (32.3 Btu) of thermal energy via theregenerator. (A negligible amount of Wout is produced at the end of thisprocess, as was described in FIGS. 11 through 17 , and above in “TDC toBDC—SD cylinder to 1st stage compressor via regenerator” through “BDC toTDC—1st stage compressor compression”.)

The 1st stage compressor now ascends towards TDC, adiabaticallycompressing the captured gas. Using the CGL calculator for an adiabaticcompression from 29.2 psi (201 kPa), 373 K, and 9.71 cu ft (275 L), to47 psi (324 kPa) and 428 K (770 R), final volume equals 6.92 cu ft,internal energy change and W-in equal 20.0 kJ (19.0 Btu).

At 47 psi and part way through the upward stroke, the 1st stage exhaustcheck valve opens and an isobaric exhaust occurs through a cooler. Perthe CGL calculator, at 373 K, the final flow volume equals 6.03 cu ft(170 L)/minute, internal energy change equals 20.1 kJ (19.0 Btu), W-inequals 8.2 kJ (7.7 Btu), and rejected heat equals 28.3 kJ (26.82 Btu).(A negligible amount of W-in is required at the end of this process, aswas described above.)

The 2nd stage compressor now descends towards BDC, compressing thecaptured gas from 47 psi, 373 K, and 6.03 cu ft to 425 K (770 R) and 74psi (510 kPa). Using the CGL calculator for an adiabatic compression,final volume equals 4.37 cu ft (123.6 L), and internal energy change andW-in equal 19.1 kJ (18.1 Btu).

Per the CGL calculator, at 74 psi, the 2nd stage exhaust check valveopens and an isobaric exhaust occurs through a cooler. At 373 K, thefinal flow volume per minute equals 3.67 cu ft (103.8 L), or 4.86 cuin/cycle (0.0796 L/cycle) @ 1,304 rpm, internal energy change equals18.5 kJ, W-in equals 8.1 kJ (7.6 Btu), and rejected heat equals 27.9 kJ.Per the ideal gas law, the H2 flow rate equals 17.07 moles/min or 0.0131moles/cycle.

An isochoric H-in now takes place, increasing the temperature from 373 Kto 466 K via the exhaust gas valved regenerator. Per the CGL calculator,pressure increases to about 94 psi. Internal energy change and H-inequal 33.7 kJ (31.9 Btu).

A second isochoric H-in now takes place, increasing the temperature toabout 547 K (984 R) via the exothermic reactor heat exchanger. Per theCGL calculator, pressure increases to about 110 psi (758 kPa). Internalenergy change and H-in equal 29.4 kJ (27.9 Btu).

Using the CGL calculator, an isobaric expansion from BDC of H2 at apressure of 110 psi (758 kPa), and a temperature of 547 K (984 R) to atemperature of 646 K (1,162 R) indicates a flow volume of 4.43 cuft/minute (125.4 L/minute) into the valved cell injector. W-out equals14.6 kJ (13.8 Btu) and H-in equals 50.6 kJ (48 Btu).

An adiabatic expansion occurs from TDC as the valved cellexpands/displaces the working fluid into the expander. Using the CGLcalculator and assuming initial parameters of 4.43 cu ft (125 L), 640.5K (1,152 R), and 110 psi (758 kPa), a final pressure of 36.5 psi (252kPa) would equal a final volume of 9.71 cu ft (275 L), a finaltemperature of 466 K (840 R). Internal energy change and W-out wouldequal 63.8 kJ (60.5 Btu).

Net W-out can be determined by a sum of these processes as shown on FIG.27 :

-   -   A. 1st stage compression adiabatic compression W-in equals 20 kJ        (19 Btu).    -   B. 1st stage compression isobaric exhaust W-in equals 8.2 kJ        (7.7 Btu).    -   C. 2nd stage adiabatic compression W-in equals 19.7 kJ (18.7        Btu).    -   D. 2nd stage isobaric exhaust W-in equals 8.1 kJ (7.6 Btu).        (Alternative is to increase adiabatic compression to ˜815 R (453        K). That corresponds to a pressure of about 92.2 psi (635.7        kPa). Volume would equal 105.7 L. Internal energy change and        W-in would equal 29.2 kJ, which is close to (19.7+8.1=) 27.9 kJ        (95.6%).)    -   E. Isochoric internal energy change and exhaust H-in equals 33.7        kJ (31.9 Btu)    -   F. Isochoric internal energy change and exothermic reactor H-in        equals 29.4 kJ (27.9 Btu).    -   G. Expander isobaric expansion W-out equals 14.6 kJ (13.8 Btu):        H-in equals 50.6 kJ (48 Btu)    -   H. Expander adiabatic expansion W-out equals 63.8 kJ (60.5 Btu).    -   I. Total W-in equals 56 kJ/min (48 Btu).    -   J. Total W-out equals 78.4 kJ/min (74.3 Btu/min).    -   K. Net W-out equals 22.4 kJ/min (21.2 Btu/min).

Examining the thermal elements of the proposed cycle, there are twoisochoric (constant volume) H-inputs and one isobaric H-input. The firstisochoric input is, of course, direct “waste heat” from the expander,and thus not included in calculating thermal efficiency. The secondisochoric input is more complex, since it is heat received from athermochemical catalytic reactor, the reactants of which are created ina separate thermochemical cycle. For the purposes of calculating theefficiency of the standalone process, it might be considered sourceheat. However, when looked at as the Benzene Battery concept's H2storage process, exothermic reactor heat may also be consideredotherwise-waste heat. It will therefore also be calculated both waysbelow.

Thermal efficiency can be determined by the ratio of net W-out minusH-in:

Total external source H-in (such as solar or geothermal) equals 50.6kJ/min (48 Btu/min).

Total exothermic H-in equals 29.4 kJ/min (27.9 Btu).

Total external source plus exothermic H-in equals 80 kJ/min (75.8Btu/min).

Total external source efficiency (exothermic heat as otherwise-wasteheat) equals 44.3%

Total source heat plus exothermic source heat efficiency equals 28%.

For such a low temperature heat engine cycle, these are good results,even for a theoretical engine.

Note that, for the cycle above, 29.4 kJ/minute of output are requiredfrom the catalytically exothermic combination of C6H6+3H2 at atemperature of about 547 K (984 R). In U.S. Pat. No. 3,225,538, Table I,chemical heat of reaction changes for C6H12<=>C6H6+3H2 are given. InTable I, chemical heat change equals approximately 52.3 kilocalories permol (219 kJ/mol, 207.5 Btu/mol) of C6H12 for both endothermic andexothermic reactions. The information given is for 1 atm (14.7 psi)constant pressure, but since heat is chemically stored, it wouldessentially be the same at any pressure or temperature driving thereaction. Thus, per minute, approximately 13.4% of a mol of C6H12 willneed to be created. That will require 13.4% of a mol of C6H6 plus 40.2%of a mol of H2 per minute. At STP, H2 (gas) has a mass of 2.02 g/mol, so0.402 moles will equal an H2 mass of about 0.812 g/minute.

Assuming the low heat of combustion, 1 g of H2 has a combustion value of˜120 kJ. Electrolysis of H2O into H2+O is about 93% efficient. Thatmeans it requires ˜130 kJ to produce 1 gram of H2. For the H2-usingprocess described above, electrolysis would require 105.6 kJ. Thus, the24% efficient thermal process described above can ideally generateapproximately 21% of the H2 required to power the cycle's exothermicheater, and the balance of the H2 required will have to come from someother source.

However, these results can be improved upon relatively easily byincreasing the amount of source temperature added per cycle. One way todo so is to raise the engine's peak temperature, as for example by theadditional H-in taking place at constant pressure (see the dottedconstant pressure H-in lines in FIG. 27 ). Such an isobaric“superheating” can be accomplished by reworking the cycle thusly;increasing the temperature of the isobarically heated H2 working fluidto the higher temperature, increasing the expansion ratio by increasingthe volume of the expander relative to the volume of the valved cell H2injector, increasing the volume of the SD cylinder to match the newvolume of the expander, increasing the volume of the 1st stagecompressor to match the increased volume of the synchronizer cylinder,and adding a third stage to the compressor/intercooler.

In one possible prototype, it will be initially assumed that the threedisplacer cylinders will be comprised of a 2.5″ diameter cylindercontaining a 2″ diameter connecting tube over a displacer piston strokeof 2.75″, or an effective piston area of 1.767 sq in and thus a cylindervolume of 4.86 cu in (0.0796 L), or 6,337 cu in/min (3.667 cu ft/min,103.8 L/min) @ 1,304 rpm.

In this design, there is no requirement for a synchronizer, since thetop displacer exhaust is isobaric. Instead, the expander exhaust isisobaric through the exhaust heat regenerator and directly into the 1ststage compressor. Since the 1st stage exhaust is likewise isobaric, thedisplace process can operate completely independent of theexpansion/compression process. Thus, there is no need to synchronize theexpander exhaust and the regenerator thermal charging processes with anSD cylinder. From above, the expander piston diameter equals 2.5″ andstroke equals 2.75″, thus piston area equals 4.91 sq in and cylindervolume equals 13.5 cu in (0.221 L). However, in this particular design,the expander has a drive rod on the top of the piston that pierces theexpander cylinder head and connects to the 1st stage compressor piston,allowing the compressor piston to be “driven” by the expander piston.

Assuming the area of the connecting rod is adjusted for by increasingthe area of the expander and compressor cylinders, the area of theexpander/1st stage compressor pistons minus the connecting rod equals13.5 cu in/cycle (0.221 L/cycle), or 288.5 L/min (10.19 cu ft/min.

An isochoric displacement takes place between the expander and the 1ststage compressor, with theoretically zero W-in or W-out. Per the CGLcalculator, an isochoric heat exchange via a regenerator willtheoretically reduce the temperature of the H2 working fluid to 373 Kand will simultaneously reduce the pressure to 17.94 psi (123.7 kPa)(Point J).

An adiabatic compression then takes place. Per the CGL calculator, anadiabatic compression in the 1st stage compressor to 29.2 psi wouldrequire compression to about 11.3 cu ft (321 L) and raise thetemperature to about 428 K. It would require W-in of about 20.9 kJ (19.8Btu) (Point K).

The 1st stage adiabatic compressor now moves towards BDC, exhausting thegas at constant pressure. The isobaric inter-cooling to 373 K and afinal volume of about 9.8 cu ft (278 L) would require about 8.3 kJ (7.9Btu) of W-in (Point A). (A negligible amount of W-in is required at theend of this process, as was described earlier.)

Per the CGL calculator, in the 2nd stage, for an adiabatic compressionfrom 29.2 psi (201 kPa), 373 K, and 9.71 cu ft (275 L), to 47 psi (324kPa) and 428 K (770 R), final volume equals 6.92 cu ft, internal energychange and W-in equal 20.0 kJ (19.0 Btu) (Point B).

At 47 psi and part way through the upward stroke, the 2nd stage exhaustcheck valve opens and an isobaric exhaust occurs through a cooler. Perthe CGL calculator, at 373 K, the final flow volume at 373 K into thethird stage compressor equals 6.03 cu ft (170 L)/minute, internal energychange equals 20.1 kJ (19.0 Btu), W-in equals 8.2 kJ (7.7 Btu), andrejected heat equals 28.3 kJ (26.82 Btu) (Point C). (A negligible amountof W-in is required at the end of this process, as was describedearlier.

The 3rd stage compressor now compresses the captured gas from 47 psi,373 K, and 6.03 cu ft to 428 K (770 R) and 75 psi (510 kPa). Using theCGL calculator for an adiabatic compression, final volume equals 4.3 cuft, and internal energy change and W-in equal 19.7 kJ (18.7 Btu) (PointD).

At 74 psi (510 kPa), the 3rd stage compressor exhaust check valve opensand an isobaric exhaust occurs through a cooler. Per the CGL calculator,at a final temperature of 373 K the final flow volume per minute intothe lower displacer equals 3.67 cu ft (103.8 L), or 4.86 cu in/cycle(0.0796 L/cycle) @1,304 rpm), internal energy change equals 18.5 kJ,W-in (of exhaust) equals 8.1 kJ (7.6 Btu), and rejected heat equals 27.9kJ. Per the ideal gas law, the H2 flow rate equals 17.07 moles/min or0.0131 moles/cycle (Point E).

An isochoric H-in now takes place, increasing the temperature from 373 Kto 466 K via the exhaust gas valved regenerator. Per the CGL calculator,H2 pressure increases to about 94 psi. Internal energy change and H-inequal 33.7 kJ (31.9 Btu) (Point F).

A second isochoric H-in now takes place, increasing the temperature toabout 547 K (984 R) via the exothermic reactor heat exchanger. Per theCGL calculator, H2 pressure increases to about 110 psi (758 kPa).Internal energy change and H-in equal 29.4 kJ (27.9 Btu) (Point G).

Using the CGL calculator, an isobaric expansion into the valved celldisplacer/injector at 110 psi to 5.1 cu ft/min (6.758 cu in/cycle @1,304 rpm) will require a temperature of about 737.4 K (1,327 R). Notethat increasing the temperature at 110 psi to 4.43 cu ft/min required aH-in of 50.6 kJ (48 Btu) and produced 36 kJ of internal energy changeand 14.6 kJ (13.8 Btu) of W-out. H-in to increase the temperature at 110psi to 737.4 K would equal an additional 49.8 kJ (47.2 Btu), internalenergy change would equal an additional 35.5 kJ (33.6 Btu), and W-outwould equal an additional 14.4 kJ (13.6 Btu). Net H-in would thus equal100.4 kJ/min (95.16 Btu/min), internal energy change would equal 71.5kJ/min (67.77 Btu/min) and net W-out would equal 29 kJ/min (27.49Btu/min) (Point L).

An adiabatic expansion from 110 psi to 22.4 psi (154 kPa) would increasevolume to about 16.05 cu ft (454.5 L) (21.25 cu in @ 1,304 rpm) with afinal temperature of 466 K. W-out would equal about 100.6 kJ (95.4 Btu)(Point M).

(Note: Clearly, that is a larger expansion than the existing prototypeexpander is capable of. Therefore, it would require either a secondaryexpander with an expansion ratio 1.635× larger or it would requirereducing the size of the three displacer cylinders and the valved cellH2 displacer/injector to 61.2% of the previously calculated size).

Net W-Out can be Determined by a Sum of these Processes:

-   -   K. 1st stage compression adiabatic compression W-in equals 20.9        kJ (19.8 Btu).    -   L. 1^(st) stage compression isobaric exhaust W-in equals 8.3 kJ        (7.9 Btu).    -   M. 2nd stage compression adiabatic compression W-in equals 20 kJ        (19 Btu).    -   N. 2nd stage compression isobaric exhaust W-in equals 8.2 kJ        (7.7 Btu).    -   O. 3rd stage adiabatic compression W-in equals 19.7 kJ (18.7        Btu).    -   P. 3rd stage isobaric exhaust W-in equals 8.1 kJ (7.6 Btu).    -   Q. Isochoric internal energy change and exhaust H-in equals 33.7        kJ (31.9 Btu)    -   R. Isochoric internal energy change and exothermic reactor H-in        equals 29.4 kJ (27.9 Btu).    -   S. Total expander isobaric expansion W-out equals 29 kJ (26.4        Btu): Total isobaric H-in equals 100.4 kJ (95.2 Btu).    -   T. Expander adiabatic expansion W-out equals 100.6 kJ (95.4        Btu).    -   U. Total W-in equals 85.2 kJ/min (80.8 Btu/min).    -   V. Total W-out equals 129.6 kJ/min (122.8 Btu/min).    -   W. Net W-out equals 44.4 kJ/min (42.1 Btu/min).

Thermal efficiency can be determined by the ratio of net W-out minusH-in:

Total external source H-in (such as solar or geothermal) equals 100.4kJ/min (95.2 Btu/min).

Total exothermic H-in equals 29.4 kJ/min (27.9 Btu/min).

Total external source plus exothermic H-in equals 129.8 kJ/min (123.0Btu/min).

Total external source efficiency (exothermic heat as otherwise-wasteheat) equals 44.2%.

Total source heat plus exothermic source heat efficiency equals 34.2%.

Note that in this approach, total ideal W-out can generate 42.0% of theH2 required to power the cycle's exothermic heater.

Note: Staged and intercooled compressors are used to approximate anisothermal compression, which is the ideal approach to compressing a gasor vapor. In the cycle above, it can be calculated that an isothermalcompression from 17.94 psi (123.7 kPa), 288.5 L/min (10.19 cu ft/min),and 373 K (671 R), to 510 kPa (74 psi) and 66.67 L/min, W-in and heatout (H-out) will both equal 50.6 kJ/min.

Net W-out equals (129.6-66.67−) 62.93 kJ/min (59.65 Btu/min).

Total external source efficiency (exothermic heat as otherwise-wasteheat) equals 62.7%.

Total source heat plus exothermic source heat efficiency equals 48.5%.

However, true isothermal compression is almost impossible to achieve ina real-world heat engine. In FIG. 27 and the calculations above, it isapproached by a three stage inter-cooled isobaric-adiabatic compressionprocess. Per the above, the requirement to use multi-staged andinter-cooled compressors created a reduction in total solar plusexothermic source heat efficiency thermal efficiency of about 30%.

A second way to increase the amount of source temperature added percycle that doesn't require increasing the peak temperature of the engineis to use a reheat followed by a “super-expansion” within a secondaryexpander. One possible prototype of a secondary reheating process, inthis instance assuming isobaric heating, is shown in FIG. 27 as a dottedline traveling from a primary expansion back to a peak temperature of646 K (1,162 R). In the proposed prototype, the previous amount ofadiabatic expansion is reduced, ending at about 60 psi (414 kPa), forexample by reducing the diameter of the primary expansion cylinder(Point N). The H2 working fluid is then pumped at constant pressure fromthe primary expander through a secondary isobaric source heater, withthe secondary heater raising the temperature once more to the peaktemperature, in this case 646 K (1,162 R). The primary expander exhaustsat constant pressure into a second valved cell injector (Point O), whichinjects the working fluid adiabatically (that is, at a constant rateduring expansion) into the #2 expander, creating a second adiabaticexpansion (Point P). Finally, the expanded gas is exhaustedisochorically from the secondary expander through theregenerator/preheater and into the 1st stage compressor, dropping thetemperature to about 373 K (671 R) (Not shown). A 1st stage adiabaticcompression takes place, followed by an isobaric exhaust through a heatexchanger, reducing the final temperature back to 373 K (Not shown). Asecond and third compression-and-inter-cooled process may then beinitiated, in this example at similar pressures and compressions to theprocess shown above. Thus, assuming similar processes as described insteps A thru H directly above:

An adiabatic expansion occurs from TDC as the valved cellexpands/displaces the working fluid into the expander (see step Jdirectly below). Initial parameters of 110 psi, 4.43 cu ft (125 L),640.5 K (1,152 R) are assumed. Using the CGL calculator, a finalpressure of 67.5 psi (462 kPa) would equal a final volume of 6.27 cu ft(178 L), a final temperature of 555 K (×R). Internal energy change andW-out would equal 30.78 kJ 29.6 Btu).

Using the CGL calculator, an isobaric expansion to a temperature of640.5 K (1,152 R) indicates a flow volume of 7.22 cu ft/minute (206L/minute) into the second valved cell injector. W-out equals 12.5 kJ(12.0 Btu) and H-in equals 43.3 kJ (41.6 Btu).

An adiabatic expansion occurs from TDC as the valved cellexpands/displaces the working fluid into the expander. Using the CGLcalculator, an adiabatic expansion to 22 psi (versus 22.4 psi (154 kPa)above) would increase volume to about 16.04 cu ft (versus 16.05 cu ft(454.5 L) above) with a final temperature of 464 K (843.5 R) versus 466K above. W-out would equal about 64.7 kJ (60.9 Btu).

Net W-out can be determined by a sum of these processes:

-   -   A. 1st stage compression adiabatic compression W-in equals 20.9        kJ (19.8 Btu).    -   B. 1st stage compression isobaric exhaust W-in equals 8.3 kJ        (7.9 Btu).    -   C. 2nd stage compression adiabatic compression W-in equals 20 kJ        (19 Btu).    -   D. 2nd stage compression isobaric exhaust W-in equals 8.2 kJ        (7.7 Btu).    -   E. 3rd stage adiabatic compression W-in equals 19.7 kJ (18.7        Btu).    -   F. 3rd stage isobaric exhaust W-in equals 8.1 kJ (7.6 Btu).    -   G. Isochoric internal energy change and exhaust H-in equals 33.7        kJ (31.9 Btu).    -   H. Isochoric internal energy change and exothermic reactor H-in        equals 29.4 kJ (27.9 Btu).    -   I. Expander #1 isobaric expansion W-out equals 14.6 kJ (13.8        Btu): H-in equals 50.6 kJ (48 Btu).    -   J. Expander #1 adiabatic expansion W-out equals 30.78 kJ (29.6        Btu).    -   K. Expander #2 isobaric expansion W-out equals 12.5 kJ (12.0        Btu): H-in equals 43.9 kJ (41.6 Btu).    -   L. Expander #2 adiabatic expansion W-out equals 64.7 kJ (60.9        Btu).    -   M. Total expander isobaric expansion W-out equals 26.1 kJ (25.9        Btu): Total isobaric H-in equals 93.9 kJ (89.6 Btu).    -   N. Total expander adiabatic W-out equals 95.5 kJ (95.5 Btu).    -   O. Total W-in equals 85.2 kJ/min (80.8 Btu/min).    -   P. Total W-out equals 121.6 kJ/min (116.4 Btu/min).    -   Q. Net W-out equals 36.4 kJ/min (35.6 Btu/min).

Examining the thermal elements of the proposed cycle, there are twoisochoric (constant volume) heat inputs and one isobaric heat input. Thefirst isochoric input is, of course, direct “waste heat” from theexpander, and thus not included in calculating thermal efficiency. Thesecond isochoric input is more complex, since it is heat received from athermochemical catalytic reactor, the reactants of which are created ina separate thermochemical cycle. For the purposes of calculating theefficiency of the standalone process, it might be considered sourceheat. However, when looked at as the Benzene Battery concept's H2storage process, exothermic reactor heat may also be consideredotherwise-waste heat. It will therefore also be calculated both waysbelow.

Thermal Efficiency can be Determined by the Ratio of Net W-Out Dividedby H-in:

Total external source H-in (such as solar or geothermal) equals 93.9kJ/min.

Total exothermic H-in equals 29.4 kJ/min.

Total external source plus exothermic H-in equals 123.3 kJ/min.

Total external source efficiency (exothermic heat as otherwise-wasteheat) equals 38.8%

Total solar plus exothermic source heat efficiency equals 29.5%.

In this model, total ideal W-out can generate 34.4% of the H2 requiredto power the cycle's exothermic heater.

Isochoric Source Heating by Internal Combustion.

As noted above, an alternative to adding constant pressure or isobaricthermal energy is to add constant volume or isochoric thermal energy(see the solid non-vertical constant volume H-in lines in FIG. 27 ). Useof combustion means to add thermal energy to a heat engine withcompressed air as its working fluid is well known. That includesrelatively simple means for rapidly injecting a fluid, vapor, or gasinto compressed air.

A valved cell can also be used to inject a gas or vapor. One advantageof using such an “injector valved cell” is that the pressure of thegas/vapor held within it only needs to equal the pressure of the gaseousand/or vaporous working fluid into which it is being injected when thecell's “transfer valve” is first opened. A process of“pressure-balancing” can then be used to aid in a rapid injectionprocess may approach constant volume.

FIGS. 30 a, 30 b, and 30 c illustrate one possible gaseous injectorvalved cell design that can intermittently connected via an exhaust ortransfer valve to an expander cylinder. Note that if injection of thequantity of gas/vapor held within the valved cell is made to exactlymatch the rate of volume expansion in the expander, an isochoric processwill occur.

A piston is situated within a valved cell cylinder and a seal separatesthe two ends of the cylinder.

The valved cell cylinder is connected on one end to an intake valve andan exhaust/transfer valve.

On the other end, a simple manifold connects to the expansion cylinder'sexpansion chamber, allowing the piston to match the pressure in theexpansion cylinder.

A cylindrical rod is attached to the piston head on the intake andexhaust side and is aligned with the axis of the plunger. The rod passesthrough the gas injecting valved cell chamber, through a seal, andthrough a small cylindrical linear bearing and a sealing ring in the“head” of the valved cell. Consequently, when both sides of the plungerare subjected to an equal gaseous/vaporous pressure, the difference inthe volume displaced by the rod will move the plunger in the directionof the gas-injecting chamber.

The cylindrical rod, having passed through the gas-injecting chamber,seal, and bearing, is connected to a low friction travel-limitingdevice, in this case a small freely rotating crankshaft. In the presentdesign, the crank throw rotates inside a horizontally-sliding bearingwhich itself is captured within a guide frame of a vertically-slidingbearing that is itself captured within a immovable guide frame. Theimmovable guide frame forces the vertically sliding bearing to travelvertically, and since the vertically sliding bearing restricts movementof the horizontally-sliding bearing to horizontal movements, therotation of the crank throw creates perfectly vertical movement of thecylindrical rod and thus of the valved cell piston.

The travel-limiting device ensures that the plunger will neverphysically contact the ends of the valved cell cylinder. It may evencreate an “accelerating injection” system as an aid in maintainingconstant volume.

During the processes of injection and refilling, there will be timeswhen the pressure on the gas-injecting side of the plunger will behigher than the pressure on the expansion chamber manifold side of theplunger. Accordingly, the plunger is constructed to be strongest on thesealing ring side, and constructed to be as light as possible on theexpansion chamber manifold side.

The piston is also constructed with sufficient length along its axis tokeep the piston sealing ring from running on the piston wall exposedintermittently on the expansion chamber manifold side of the cylinder.If necessary, the portion of the valved cell injector cylinder wall inwhich the piston sealing ring sits is cooled.

The gas intake valve may be a simple poppet-type check valve that sealsagainst higher pressure on the injection chamber side. It is biased toreturn to closed, for example by a return spring.

The exhaust or transfer valve is constructed similarly to the CCVCprototype transfer valve. It is opened primarily by pressureequalization across the valve head at the end of the expander exhauststroke. The pressure equalization is created, as in the CCVC prototype,by a slightly early closure of the expander exhaust valve that capturesa small amount of remnant gas in the limited space between the expansionpiston and the expansion head, thus allowing pressurization of saidremnant gas to equalize pressure across the transfer valve. Whenpressure is thus equalized, some means, such as a spring bias, may beused to easily and quickly open the transfer valve, thus connecting theexpander cylinder to the previously-charged gas-injecting chamber.

Upon movement of the transfer valve towards open, the force differentialacross the valved cell piston (created by the cylindrical rod) willbegin to inject the gas into the expander combustion chamber.

If there is a single gaseous fuel injector such as H2 injecting into anoxidizing environment, the fuel and oxidizer instantly begin to mix andmay be instantly combusted, instantly driving up the pressure in thecombustion chamber, the valved cell injector chamber, and the manifoldconnecting to the other side of the valved cell piston. That in turninstantly drives the valved cell piston to inject all the contents ofthe injector chamber into the combustion chamber.

If there are two gaseous injectors, as for example a gaseous O2 injectorand a gaseous H2 injector, then both injector exhaust or transfer valvesopen simultaneously by the same process. As a result, O2 and H2instantly begin to mix and instantly combust, instantly driving up thepressure in the combustion chamber, the H2 valved cell chamber, the O2valved cell chamber, the expander cylinder, and the manifold connectingto the O2 valved cell plunger. That in turn instantly drives the O2valved cell plunger to inject all the O2.

Referring to the above, FIG. 27 , and the CGL calculator, at 74 psi and373 K (Point E), the final flow volume per minute equals 3.67 cu ft(103.8 L), or 4.86 cu in/cycle (0.0796 L/cycle) @ 1,304 rpm, internalenergy change equals 18.5 kJ, W-in equals 8.1 kJ (7.6 Btu), and rejectedheat equals 27.9 kJ. Per the ideal gas law, the H2 flow rate equals17.07 moles/min or 0.0131 moles/cycle.

Assuming an isochoric thermal input to 555.6K, final pressure wouldequal 110.2 psi, and recycled H-in would equal 63.82 kJ (Point G).

Assuming an isochoric thermal input to the peak temperature of 950 K(1,710 R, 677 deg C., 1,250 deg F.) (Point Q) shown in FIG. 27 , an H2flow rate of 17.07 moles/min or 0.0131 moles/cycle, and a volume of 3.67cu ft/min (103.8 L/min), pressure equals 1300 kPa (188.5 psi), andsource H-in would equal 138 kJ. An adiabatic expansion to 555 K and 13.8cu ft (390.6 L) would equal about 29.3 psi (202 kPa). Total change ininternal energy change and W-out equals 138 kJ/min.

The instant the gas/vapor injector valved cell connected to the expanderare emptied, an adiabatic expansion then occurs within the mainexpander. Note that, at expander BDC, continued expansion may followinto a lower pressure, lower temperature, uncooled and non-lubricatedsecondary expander.

When the pressure in the gas/vapor injector valved cell is equal to thefeed pressure of the gas/vapor sources (in this case, at 110 psi,occurring when the adiabatic expansion drops to about 1,580 R or 877.7K), the gas/vapor injector valved cell exhaust or transfer valve isclosed. As soon as said injector exhaust valve is closed, the gas/vaporintake valve is opened.

When pressure in the expander drops below the feed pressure of thegas/vapor sources, pressure is reduced in the pressure-equalizingdisplacer system, and the gas/vapor injector valved cell cylinders areautomatically refilled.

When the gas/vapor injector valved cells are fully charged, thegas/vapor injector valved cell intake valve is closed, completing thecycle and preparing for the next cycle.

If the gas/vapor is derived from a liquid, excess heat from the heatengine can be used to preheat the gas/vapor to a high pressure with verylittle W-in. Note that there is then no requirement for a gascompressor.

In the instance that an H2+O2 combustion process occurred, followingadiabatic expansion and exhaust through the preheater, the working fluidmay be cooled sufficiently for H2O to be easily separated from thenon-combusted working fluid gas/vapor. The remnant working fluid maythen be recycled through the compression system, as described above

Continuing on, the exhaust from the expander (at about 1,000 R (555 K)in this instance) can either be isobaric or isochoric.

For an isobaric exhaust, the usual approach (and the approach used inthe original CCVC prototype) would be to “capture” thermal energy bymeans of a counterflow heat exchanger. As a result:

-   -   (1) The two counter flowing streams of fluid are required to        each have their own “containers”, and heat is only able to        transfer by conduction through the walls of those containers        (usually made up of many small tubes in direct physical contact        with one another). This results in relatively poor heat transfer        over time.    -   (2) In order to give the heat transfer process more time to take        place, the tubes are generally quite long, creating a large        amount of internal volume, thus resulting in low changes in        temperature and pressure over a short distance, should that be        required by one or the other fluid streams.    -   (3) Because of the length of the heat exchanger and, in a heat        engine, the requirement for the receiving fluid to be at a much        higher pressure and thus the receiving fluid container to be        stronger than the thermal charging fluid, a great deal of mass        must be heated/cooled.    -   (4) Since the exhaust process is isobaric, the work of exhaust        is relatively high.    -   (5) Assuming an isochoric heat absorption process, an isobaric        exhaust will theoretically contain more thermal energy than the        isochoric process can use, and thus may represent waste energy.

For an isochoric regeneration, the advantages are:

-   -   (1) Since the gases pass through a “thermal sponge”, the        internal masses required of the heat exchanger are greatly        reduced, since heat is given off in one flow and taken in in the        opposite flow. This results in a highly efficient heat transfer        process.    -   (2) The internal volumes are greatly reduced, resulting in much        higher changes in temperature and pressure for a distance        traveled by the fluids.    -   (3) No W-in is required in exhausting isochorically from the        exhauster displacer volume, and, since thermal energy is        removed, the receiver displacer volume's exhaust will be at a        much lower pressure and temperature, thus requiring less work        overall.    -   (4) Assuming a match in mass displaced in both directions, there        is exactly as much thermal energy charging the regenerator (at a        lower pressure) as is required by an isochorically-displaced        (higher pressure) gas that will be absorbing the thermal energy.

An isochoric expansion to about 474 L/min (16.73 cu ft/min), 373.5 K(670 R), and 112.4 kPa (16.3 psi) would generate 64.0 kJ of heat. In anideal cycle, an isothermal compression to 373.5 K (670 R) and 74 psi(510 kPa) would reduce volume to 104.4 L/min (3.68 cu ft/min). W-in andH-out would equal 80.5 kJ. Since exactly as much thermal energy would beavailable in the exhaust at constant volume as in the regeneration intothe compressed working fluid at constant volume, then it is easier tocalculate the potential ideal thermal efficiency if the exhaust throughthe regenerator were at constant volume.

An isochoric exhaust process proceeding from 24.1 psi (166.2 kPa) at atemperature of 555 K (1,000 R) and a volume of 474 L/min (16.73 cuft/min) to a pressure of 112.4 kPa (16.3 psi) and a temperature of 373 K(670 R) would have an internal energy change and a heat rejection of64.0 kJ

For an isothermal compression, an isothermal compression from 474 L/min(16.73 cu ft/min), a pressure of 112.4 kPa (16.3 psi), and a temperatureof 373 K (670 R) to a final pressure of 510 kPa (74 psi) and a finalvolume of 104.4 L/min (3.69 cu ft/min) would equal H-out and W-in equalto 80.5 kJ. Total work generated equals 57.3 kJ/min. Thermal efficiencywould thus equal total W-out divided by total H-in or 41.5%.

However, as stated above, a three stage inter-cooled isobaric-adiabaticcompression process will require about 85 kJ/min, reducing overall W-outto 53 kJ/min. Thus, the theoretical thermal efficiency of the aboveprocess would equal about 38%.

Since the peak temperature is 950 K and the sink temperature equals 373K, theoretical thermal efficiency equals (T1−T2)/T1 or 60.7%, thetheoretical thermal efficiency equals 62.5% of Carnot.

Isochoric Source Heating+Exothermic Preheating

It is quite possible to use multiple isochoric regeneration as isdescribed above to replace some high grade source heat with medium gradesource heat. In FIG. 22 , at 1 atm, for a 90% conversion of H2+C6H6 intoC6H12, thermal output would equal about 650 K (1,170 R). Note that, inU.S. Pat. No. 3,225,538, Table I, chemical heat change equalsapproximately 52.3 kilocalories per mol (219 kJ/mol, 207.5 Btu/mol) ofC6H12 for both endothermic and exothermic reactions. At a 95% conversionrate, that would equal 197 kJ per mol. Thus, per minute, approximately16.8% of a mol of C6H12 will need to be created. That will require 16.8%of a mol of C6H6 plus 50% of a mol of H2 per minute. At STP, H2 (gas)has a mass of 2.02 g/mol, so 0.5 moles will equal an H2 mass of about 1g/minute of H2.

Per the CGL calculator, to isochorically raise the temperature of themix from 555 K to 650 K would increase the pressure to 888.8 kPa (128.9psi) and require 33 kJ/min, decreasing the source heat required to 105kJ. That in turn increases the thermal efficiency to 50.4%, andincreases the percentage of Carnot to 83%.

Finally, since W-out equals 53 kJ and electrolysis is 93% efficient,this model can produce 49 kJ worth of H2 per minute. Since, from above,it requires 130 kJ to produce 1 gram of H2, this model can produce about0.38 grams of H2/minute, in this model, total ideal W-out can generate38% of the H2 required to drive the exothermic reaction.

Isochoric Source Heating by Internal Combustion in Compressed Air.

One possible prototype would involve injecting pressurized H2 intocompressed and preheated air in an “open” cycle process (see oneproposed gaseous injector design above under the “Isochoric sourceheating by internal combustion” heading). Assuming the “used” air isexhausted following the heat content being removed to preheat a newcharge of compressed air, then a fuel, such as pressurized H2, can beinjected with essentially the same potential efficiencies determinedabove.

Assumptions:

Use of compressed air as the working fluid at 74 psi (510 kPa) and 670 R(372 K); i.e., open cycle isochoric combustion.

Use of compressed H2 as the fuel.

Use of the existing prototype expansion ratio of 1:2.777; 4.86 cuin/cycle (0.0796 L/cycle) displaced into 13.5 cu in/cycle (0.221L/cycle); 3.67 cu ft/min (103.8 L/min) displaced into 10.18 cu ft/min(288.2 L/min).

The upper displacer cylinder is used as the valved cell.

Two intercooled compressions totaling approximately 50 Btu/min; anadiabatic expansion totaling approximately 75 Btu/min; total theoreticalW-out totaling approximately 25 Btu/min (26.4 kJ) or about 0.6 HP/hr(0.44 kWh).

Peak temperature of about 1,180 R (655.6 K); expander exhausttemperature of about 840 R (466.7 K); synchronizer displacer exhausttemperature of about 670 R (372 K).

Peak pressure would approach 120 psi (827 kPa); expander exhaustpressure would equal about 30 psi; synchronizer displacer exhaustpressure would equal about 25 psi (note that a small turbocharger coulduse this exhaust energy to “boost” the input stream to the 1st stagecompressor), improving thermal efficiency).

H-in equal to about 75 Btu/min or 0.0575 Btu/cycle (60.7 J/cycle);assuming the low heat of combustion of 1 g of H2 or 120 kJ, the mass ofthe injected H2 would equal approximately half a mg/cycle (30.3mg/minute, 1.8 g/hour, at STP, H2 (gas) has a mass of 2.02 g/mol, andthe molar mass of H2 injected per cycle would equal 0.00025 moles;Pressure at the end of isochoric waste heat regeneration would equalabout 80 psi (551.6 kPa) and temperature would equal about 840 R (466.7K); per the ideal gas calculator, at 0.00025 moles, 466.7 K, and 551.6kPa, injector volume per cycle would equal 0.00176 L (0.107 cu in).Assuming a 0.75″ diameter injector cylinder, stroke would equal about0.25″.

Overall theoretical thermal efficiency (assuming no turbocharger) equals25 Btu W-out divided by 75 Btu H-in or 33.3% (about equal to the maximumefficiency of a typical gasoline-burning engine). Ideal Carnotefficiency or T1/T2)/T1 equals 43.2%. Percentage of ideal Carnotefficiency thus equals 77.1%.

Isochoric Heating with an Isobaric Heat Source

An isobaric valved regenerator or STREP is proposed in FIGS. 1 through 4. Note that, for example, the thermal output of an exothermic reactorcan be easily and efficiently absorbed by a coolant such as H2recirculated at constant pressure by an isobaric STREP.

As mentioned above, a different kind of STREP was proposed in FIGS. 11through 17 , where the SD cylinder received the output from the expanderand the “receiver cylinder” served double-duty as the 1st stagecompressor. Unlike the STREP in FIGS. 1 through 4 , which exchangethermal energy isobarically or at constant pressure, the STREP in FIGS.11 through 17 exchanged thermal energy isochorically or at constantvolume.

It is likewise perfectly feasible to thermally charge a STREP with anisobaric gas flow, then “switch” the regenerator to isochorically removesome or all of the thermal charge thus deposited. FIGS. 32 and 42illustrate what such a mixed isobaric/isochoric STREP might look like.

In the relatively simple cycle shown in FIG. 33 , both solar energy andexothermic reactor energy are shown being absorbed by constant pressure“carrier fluids”, capturing thermal energy at the peak temperature ofboth processes. In both examples, a mixed isobaric/isochoric STREP isused to then efficiently transfer much of that thermal energy intoisochoric processes that are thus “driven” to the peak temperatures ofthe heat sources. A valved cell displacer/injector and adiabaticexpander then generates work, and the exhaust energy is finally capturedas well in an isochoric STREP. The result, as shown in FIG. 33 , is tocreate a series of four isochoric thermal inputs: First, an isochoriccapture is followed by an isochoric regeneration of waste exhaust heatfrom the engine by using a STREP similar to the one shown in FIGS. 11through 17 ; second, an isobaric capture and an isochoric regenerationof exothermic reactor heat; and finally, an isobaric capture and anisochoric regeneration of high temperature source energy such as solarenergy.

FIG. 42 illustrates a process whereby a constant-pressurize gas, havingabsorbed thermal energy from some process, can, by “matching” pressuresduring a charging cycle with pressures on the isochoric side of aregenerator, in turn be used as the carrier of said externally appliedthermal energy, passing it cyclically through a valved regenerator.

Note that FIG. 42 shows a much larger cylinder on the lower left than onthe lower right, and a much smaller cylinder on the upper left. If thelarger lower cylinder is at high temperature and low pressure, then itsvolume will be reduced as it deposits thermal energy into theregenerator, reducing the volume being displaced into the upper leftcylinder. FIG. 42 is thus illustrating the impact on an isobaricdisplacement of removing thermal energy over the course of a givenstroke distance. In this case, the higher pressure thermally receivingfluid will be in the cylinder on the lower right, which will receivefluid at constant pressure as the pistons move downward.

In another use of the STREP in FIG. 42 , a reactant mix, for examplecomposed of one mol of C6H6 and three moles of H2, is passed through theregenerator at constant pressure and temperature, delivering thermalenergy to a regenerator that is also a catalytic reaction chamber. As aresult, the number of moles of product is potentially reduced to aquarter of the moles within the reactant thus created. And since neitherpressure nor temperature changed, volume is correspondingly reduced toone quarter as well.

As stated earlier, a high temperature but low pressure exhaust fluid atconstant pressure can be passed through a counterflow regenerator-typeheat exchanger, thermally “charging” the regenerator. The regeneratorcan then be raised in pressure, in this case by a semi-adiabaticcompression of remnant product in the large cylinder by an early closureof the regenerator's exhaust valve. In this instance, a separate streamof counter-flowing fluid at the higher pressure but at low temperaturecan then enter the regenerator through an intake valve and flowisobarically through the regenerator into the lower right cylinder, thusisobarically absorb the thermal energy deposited by the earlier lowpressure flow. Note that it could also be a higher pressure isochoricabsorption process. Finally, the high pressure in the regenerator can bereduced to that of the low pressure stream and once more be used to“charge” the regenerator, in this case by early closure of the upperleft small cylinder exhaust valve causing a semi-adiabatic compressionof remnant product there, followed by a re-expansion of remnant highpressure fluid in the regenerator back into the low pressure displacerat the beginning of its intake stroke, followed by low pressure fluidflow out of the large lower cylinder into into the small upper cylinder.

In other words, it is perfectly feasible to thermally charge a valvedregenerator with a gas at isochorically, then “switch” the regeneratorto isobarically add or remove some or all of the thermal charge thusdeposited, or vice versa.

Work-In Requirements of a C6H6+3H2 Exothermic Heat Generator.

An important requirement is that C6H6+3H2, also called an “exothermicfluid”, be made available in a state that is capable of generating therequired temperature for conversion into C6H12, also called an“endothermic fluid”, in this case exothermic heat produced at atemperature 547 K (984 R). Per FIG. 22, that would require a pressure ofapproximately 2 atmospheres (30 psi, 206 kPa).

When a heat engine both produces W-out and useful thermal energy, it istermed a CHP or Combined Heat and Power process. In U.S. patentapplication Ser. Nos. 17/746,848 and 18/095,463, it is proposed thatconversion of C6H6+3H2 into C6H12 can be done in conjunction with oreven within the confines of heat engines that generate useful thermalenergy, W-out, or a combination of both useful thermal energy and W-out.For generating W-out only, the process is termed a Bland/Ewing CombinedCycle or B/E-CC process. For generating both W-out and useful thermalenergy, the process is termed a Bland/Ewing Combined Heat and Power orB/E-CHP process.

In the B/E-CHP process, when some smaller portion of a larger amount ofexothermic fluid needs to be converted to endothermic fluid, the heatthus generated can drive an engine used to supply power for the processof converting the whole of the endothermic fluid. In U.S. patentapplication Ser. No. 18/095,463, FIG. 30 and FIG. 35 , an exothermic“production” B/E-CHP Cycle is also proposed. It is estimated to generatea small amount of excess W-out while also converting on the order of 75%of the exothermic fluid into useful thermal energy. Note that this cyclewas not “optimized” to take advantage of the STREP heat exchangeprocess.

One proposed mechanism for optimizing a B/E-CHP process through use of amodified STREP would involve:

-   -   1. Physically connecting two equal volume displacers with        opposing cyclical intake and exhaust strokes via a regenerator;    -   2. adding an intake and an exhaust valve to each displacer,        where the exhaust of the lower temperature displacer connects to        the cold side of the regenerator and the intake of the higher        temperature displacer connects to the hot side of the        regenerator;    -   3. Adding an additional intake valve and exhaust valve to the        regenerator itself, where the regenerator intake valve connects        to the hot side of the regenerator and the regenerator exhaust        valve connects to the cold side of the regenerator;    -   4. The regenerator intake valve would be connected to the        adiabatic high temperature, high pressure side of a counterflow        recuperator, which would receive thermal input from some heat        source, for example the high temperature, low pressure exhaust        from a heat engine;    -   5. The regenerator exhaust valve would be connected to the        adiabatic low temperature, high-pressure side of said        counter-flow recuperator, which would connect to the low        temperature side of said counterflow recuperator.    -   6. A double-acting intermittent and cyclical pump located near        to the regenerator exhaust valve receives cold constant pressure        working fluid from the cold side of the regenerator and pumps it        into the system entering the cold side of the counterflow        recuperator. This cyclical pump is timed to intake fluid into        one side of the double-acting piston and simultaneously exhaust        fluid from the opposite side when the regenerator intake and        exhaust valves are open. On the following cyclical opening of        the regenerator intake and exhaust valves, the opposite side        takes in a “charge” of cold fluid while exhausting the previous        “charge” with the double-acting piston. This double-acting        intermittent pump mechanism can, for example, be “timed” to        operate through use of a “geneva mechanism”, in a manner that is        common practice.

The cyclical process would take place in the following manner:

-   -   a. As the higher temperature displacer reaches its maximum        volume and the lower temperature displacer reaches its minimum        volume, the displaced contents will reach its maximum pressure        and temperature throughout the connected displacers and the        regenerator, and the higher temperature displacer intake valve        is closed.    -   b. The higher temperature displacer exhaust valve, which may be        the expander transfer valve, is instantaneously opened, allowing        the fluid in the higher temperature displacer to exhaust. (Note:        The exhaust may be to a second isochoric temperature input        system, or may be the exhaust to an isobaric temperature input        system. However, for this example, it will be assumed to exhaust        directly into an expander.) Simultaneously, the lower        temperature displacer begins to move away from minimum volume,        and the lower temperature displacer exhaust valve is held open        by some force, dropping the pressure in the regenerator and the        lower temperature displacer.    -   c. The pressure in the regenerator and lower temperature        displacer drops to approximately the pressure in the        counter-flow recuperator system. Simultaneously, the exhaust        valve for the lower temperature displacer instantaneously closes        and the regenerator intake and exhaust valves open.    -   d. A measured quantity of the adiabatic fluid proceeding from        the recuperator at high pressure and high temperature is then        “pumped” through the regenerator, thermally charging it.    -   e. As the higher temperature displacer approaches its minimum        volume and the lower temperature displacer approaches its        maximum volume, the higher temperature displacer and expander        contents will reach its minimum pressure and temperature        throughout the connecting manifold, and the higher temperature        displacer exhaust valve is closed.    -   f. Continued travel of the higher temperature displacer will now        raise the pressure of any remnant fluid until it reaches        approximately the pressure in the regenerator. As the higher        temperature displacer reaches its minimum volume and the lower        temperature displacer reaches its maximum volume, at which point        the higher temperature displacer intake valve will open.        Simultaneously, the regenerator intake and exhaust valves are        closed, and the lower temperature displacer intake valve is        opened. (Note: The exhaust may be to a second isochoric        temperature input system, or may be to an isobaric temperature        input system. In that case, the higher temperature displacer        intake valve is not opened until some re-expansion of remnant        fluid within the higher temperature displacer occurs, equalizing        pressure between it and the regenerator, at which time the        higher temperature displacer intake valve will be opened.)    -   g. A purely isochoric displacement occurs. The isochoric        displacement through the regenerator then raises both the        temperature and the pressure of the displaced working fluid        until the upper displacer has completed its expansion and the        lower displacer has completed its exhaust, and the cycle begins        again.

A C6H12 Dissociation-Pressurized H2 Gas Generator (See FIG. 38 and FIG.41 )

For 0.4536 kg C6H12 converted 100%, the yield is 0.4210 kg of C6H6 and0.0326 kg of H2.

The vapor molar heat capacity of C6H12 is 105 J/(mol K), or a vapormolar heat capacity of 1.25 kJ/kg/(K).

The vapor molar heat capacity of C6H6 is 82.4 J/(mol K), or a vapormolar heat capacity of 1.05 kJ/kg/(K).

The molar heat capacity of H2 is 28.84 J/(mol K) (6.89 cal, 0.0273 Btu),or 14.27 J/gram/(K), or 14.28 kJ/kg/(K). (For 1 lb (0.454 kg), molarheat capacity equals 6.48 kJ/(K).)

The total molar heat capacity of one mol K of C6H6 plus 3 moles of H2equals 168.92 J/(K).

C6H12 boils at 1 atm and 353.9 K (637.0° R). C6H12 has a standard heatof vaporization requirement of 32 kJ/mol/(K), or 380 kJ/kg.

C6H6 boils at 1 atm and 353.2 K (635.8° R). C6H6 has a standard heat ofvaporization requirement of 33.9 kJ/mol/(K), or 433 kJ/kg.

Per FIG. 22 , at 950 K (1,710 R, 677 deg C., 1,250 deg F.), anendothermic reaction would require a pressure of about 5.25 atmospheres(77.2 psi, 532 kPa). It will be assumed that 1 mol of C6H12 (theproduct) will be converted per stroke to 1 mol of C6H6 plus 3 mols of H2(the reactant) with a total molar heat capacity of 168.92 J/degree K.The total molar heat content of 1 mol of the reactant will thus equal160.474 kJ.

As noted above, U.S. patent application Ser. No. 18/095,463, use of anExothermic Reactor Exhaust Compressor (EREC) is proposed to assist inthe vaporization of C6H6 by a counter flowing exchange of heat withcondensing higher pressure C6H12. An Endothermic Reactor ExhaustCompressor (ENREC) may likewise be used to permit the condensation of 1mol of higher pressure C6H6 to supply all of the thermal energy requiredto vaporize 1 mol of C6H12. Note that C6H6 and C6H12 boil atapproximately the same temperature and have approximately the samestandard heat of vaporization requirement. Since the total molar heatcapacity of C6H12 at 950 K equals 99.750 kJ, only 62.2% of the totalmolar heat content of the reactant is required to preheat C6H12 from atemperature just above vaporization, estimated at 423 K, to thetemperature required for the endothermic reaction at 950 K and 532 kPa.The difference in temperature being 418 K, total heat available wouldequal 70,608 kJ, heat transferred would equal 43.919 kJ, and remainingheat would equal 26.689 kJ.

It is therefore possible to separate the exothermic fluid exiting thereactor into a 62% stream and a 38% streams, the 62% stream being usedto vaporize the endothermic fluid. The reactant mix will now beexhausted through two different heat exchangers. In one possible usecase, the smaller fraction will pass through heat exchanger #2 andpreheat H2 returning from the gas/liquid separator, as will be shown.The 62% stream will enter the main endothermic reactor preheater,preheating the vaporous endothermic fluid to the endothermic reactortemperature and simultaneously cooling the reactant mix to just aboveC6H6 condensation temperature. At which time the two reactant mixstreams will be recombined, as will be shown. Note that this may be doneat constant pressure or at constant volume.

The combined streams now enter the exothermic mix condenser/endothermicmix vaporizor/cooler, which is located between the liquid C6H12 pump andthe (endothermic) ENREC compressor. The combined flow of higher pressurereactant will supply the thermal energy to preheat and vaporize thelower pressure liquid C6H12, after which the C6H12 vapor will becompressed by the ENREC to about the pressure of the reactant, in thiscase to 5.25 atm. Finally, as flow continues, the reactant will enterthe cooler and be cooled completely, thus separating the liquid C6H6 andremnant C6H12 from the H2 gas. The C6H6 and C6H12 can then be furtherseparated, as by use of a centrifuge.

Note, however, that there is a huge volume difference between thereactant and the product at any given temperature and pressure, sincewhat was a single mol is now 4 moles. This is somewhat analogous to thehigh temperature low pressure exhaust gas from a combustion engine beingused to preheat low temperature high pressure working fluid for theengine where, unlike is shown in FIGS. 1 through 4 , a STREP exchangingthermal energy between those two adiabatic fluid flows would need tohave a much larger low pressure cylinder and a much smaller highpressure cylinder, similar to that shown in FIG. 34 . Of course, in thecase of the vaporous endothermic fluid and the vaporous exothermicfluid, the pressures would be similar, but the volumes would still behugely different, especially so considering the possession of bothtemperature and molecular differences.

The pure H2 gas at 5.24 atm is now free to be used. There are severalpossibilities:

As suggested in U.S. patent application Ser. No. 18/197,092, FIG. 27 ,the separated 5.25 atm H2 can be passed back through a H2/reactantisobaric recuperator-style heat exchanger/preheater. Alternatively, theH2 can be passed back through an H2/reactant isobaric STREP preheater,either a constant pressure version as described in FIGS. 1 through 4 , aconstant volume version as used in FIGS. 11 through 17 , or a mixedversion as described herein. See FIGS. 35 through 41 illustrating onepossible Bland/Ewing composite cycle to which the STREP heat exchangeprocess can be applied.

FIG. 35 shows all the paths in FIGS. 36, 37, and 38 as a tracing overFIG. 18 . FIGS. 36 through 38 show the working fluid pathways based onFIG. 18 for a work producing cycle and a refrigeration cycle. FIG. 36shows the paths for a low pressure and a higher pressure exothermichalf-cycle that produces endothermic fluid (in this instance C6H12).FIG. 37 shows the H2 compression path for both an exothermic andendothermic work-producing cycle and the H2 path for the exothermicwork-producing cycle. FIG. 38 shows the endothermic half-cycle path fora refrigeration cycle.

FIGS. 36 through 38 are labelled. The labels can generally be applied toFIGS. 39, 40, and 41 , as shown. FIG. 39 is copied from FIG. 25 andshows a simple exothermic heat generation system, said produced heatwhich can then help power or completely power an exothermic half-cycleengine or otherwise produce useful heat.

FIG. 40 is copied from U.S. patent application Ser. No. 18/197,092, FIG.2 , and shows a higher pressure work-producing cycle such as is shown inFIG. 37 . FIG. 41 shows FIG. 40 reconfigured as a refrigeration cyclesuch as is shown in FIG. 38 . (Note: FIG. 40 indicates an ERECcompressor which has been relabeled an ENREC compressor. To avoidconfusion, FIG. 41 also indicates an EREC compressor that is actually anENREC compressor.)

FIG. 36 , in addition to show a paths for a low pressure and a path fora higher pressure exothermic half-cycle that produces endothermic fluid,is also comparing an alternative approach to using an EREC to raisepressure following vaporization. The thin dotted line between A and D orA′ and D′ represents the adiabatic compression of an EREC to the higherpressure at which the catalytic reaction (exothermic in this instance)will take place. The line from D to B or D′ to B′ represents an isobaricheating to the temperature of the reactor, while the line from B to C orB′ to C′ represents an isobaric regeneration of the heat from theexhausting product into the cold reactant.

However, it is also possible to raise the pressure of the reactant byusing an isochoric rather than an isobaric process such as via a STREPheat exchange process, and use the exhausting product to supply thatthermal energy, although it would not be as thermally efficient. Notethat the product would still take the line from B to C or from B′ to C′,since it is desirable that the product be at the higher pressure inorder to supply the heat of condensation of the liquid constituent ofthe product to accomplish the vaporization of the liquid or solidreactant.

In essence, an isochoric STREP process can be seen as accomplishing akind of “thermal isochoric compression” as opposed to anadiabatic/isentropic compression.

FIG. 36 further shows a thin rising curved line between C and D′. Thisthin line represents the impact of rising pressure on the point at whicha liquid converts to a vapor. As shown, at about 5.25 atm, thetemperature required to, in this instance, vaporize C6H12, has increasedto about 750 R (416 K, 290 deg F., 144 deg C.). Note that, at thatpressure, using an isochoric STREP process can essentially only functionabove that temperature. However, recall that pressure is supplied to theliquid or solid reactant, avoiding the requirement to compress a vapor.While that might seem to “save” considerable W-in, an isobaric systemalso takes W-out as thermal energy is added and returns that work asthermal energy is removed, so the net efficiency gain favors theisobaric STREP process.

Recall that ideally mol count would equal 3 mols of H2, for a totalmolar heat capacity of 86.52 J/degree K. Recall that the remaining heatavailable equals 26.7 kJ. Therefore, if passed through a purely isobaricreheater, assuming a 100% efficient heat transfer, the 3 mols of H2 at418 K can be raised 308.5 K, to 726.5 K.

Alternatively, the 5.25 atm H2 can be passed back through an H2/reactantisochoric STREP or mixed isobaric/isochoric STREP (FIG. 32 ). Per theideal gas law calculator, 1 mol of H2 at 423K and 532 kPa equals 6.61 L.For 3 mols, volume equals 19.8 L. Per the CGL calculator, for anisochoric thermal input of 26.7 kJ, the final pressure would equal 1,079kPa and the final temperature would equal 858 K. Thus pressure would beincreased by 547 kPa, or essentially double the pressure of the isobaricSTREP or the typical isobaric recuperator, and temperature would beincreased by 131.5 K to nearly the temperature of endothermicdissociation.

Note that the volume of the H2 cylinder would be much closer to thevolume of the C6H6+H2 exothermic fluid mix cylinder at a similartemperature and pressure. No calculations have been attempted on thispossible approach.

A second interesting alternative for using the H2 is as a refrigerant.Having cooled the C6H12 and H2 below C6H12 condensation temperature(estimated at 670 R (372 K, 99 deg C., 210 deg F.) and thus separatedout the H2, the 5.25 atm and 3 moles of H2 can then be further chilledto ambient temperature and expanded to generate cold. Per the ideal gascalculator, volume prior to expansion would equals 17.44 L. Per the CGLcalculator, expanding 56.77 L of H2 at 372 K to from 5.25 atm to 1 atmwould decrease the temperature of the H2 to 230.7 K (−42.4 deg C., −44.4deg F.) and would generate 8.684 kJ or work. Assuming a final isobaricexhaust at 1 atm, no exhaust W-in or W-out is required. Note that the26.7 kJ of thermal energy at 950 K is still available.

The cooled H2 gas at 5.25 atm can be stored for later use, and anyexcess latent heat can be used for CHP and/or CC purposes.

The cooled and expanded H2 gas can be fed to a low pressure fuel cell,generating electricity at very high efficiency.

The cooled gas can be injected with no additional compression requiredinto the cooled H2 working fluid following the final compression of anSD CVCC H2+O2 combustion engine. That is, it and can be used as “makeupH2” to replace the combusted H2. Note that a gas compressor, preferablymulti-staged and inter-cooled, is still required for the majority ofcycling working fluid. Recall that, in order to keep peak combustiontemperature down to a sustainable level, there must be a large quantityof non-combusted working fluid relative to the combusted working fluid.

If H2 is used as the working fluid, a simple injection ofpre-pressurized O2 into pressurized and preheated H2, for example bycyclically injecting the O2 via a “displacer” valved cell (see proposedgaseous injector, FIGS. 30 a, 30 b, and 30 c ), the injected O2 cansupply the required heat of combustion to drive the cycle. Note that theoriginal pressure of the gas taken into such a displacer valved cellonly needs to equal the pressure of the gaseous or gaseous and/orvaporous working fluid in the H2 valved cell prior to combustion.Finally, note that the H2O thus produced in the resulting working fluidmix can easily be condensed and removed at the end of the cycle. Infact, the removal of the H2 captured in the newly-produced H2O is why“makeup H2” is required.

If air is used as the working fluid, for example by cyclically injectingthe H2 via a “displacer” valved cell, the air plus water/steam mix canbe “dumped” each cycle. Note that an H2 displacer valved cell would bemuch larger than an O2 displacer valved cell.

C6H12 as a Regenerator Thermal Charging Fluid in a C6H12 ProductionSystem.

FIGS. 1 through 4 illustrate a means for accomplishing an isobariccyclical regeneration. As suggested above, such an isobaric cyclicalregeneration can utilize C6H12 as a “regenerator thermal charging fluid”in the conversion of C6H6+H2 to C6H12. A C6H12 regenerator chargingfluid system would comprise the steps of:

-   -   a. Relatively cold C6H6 vapor at constant pressure (P1) and        temperature (T1) would intermittently and cyclically pass        through regenerator intake valve (In1), passing a stream of        working fluid from the “cold” side of a regenerator (A) to the        “hot” side, pass out through a port on the “hot” side of the        regenerator, and pass through the intake valve (In2) and into a        C6H6 receiver mechanism, such as a piston-and-cylinder        arrangement (B). The result is to increase the temperature of        the C6H6 at constant pressure, and simultaneously remove some or        all of the regenerator's stored thermal energy.    -   b. Simultaneously, a second receiver mechanism, such as a        piston-and-cylinder arrangement (C), would isobarically receive,        at a similar pressure of P1 and a temperature of T2, a charge        through its intake valve (In3) of relatively hot C6H12 vapor        exiting an endothermic catalytic reactor.

Note, importantly, that the volume of the two receiver mechanisms may ormay not be equal, depending on various factors such as relativetemperatures at the end of this stroke or the chemical nature of the thetwo streams. FIG. 34 graphically indicates a case where, for example, acharge of vaporous C6H6+H2 is used to preheat a charge of vaporousC6H12, in this case in order to preheat C6H12 which is about to enter anendothermic reactor and be converted to a new charge of C6H6+H2.

-   -   c. Both receiver mechanisms, having filled completely, would now        reverse direction. Simultaneously, valves In1, In2, and In3 will        be closed, and valves Ex1, Ex2, and Ex3 will open. Valve Ex1        exhausts the heated C6H6, valve Ex2 exhausts the hot C6H12 and        acts as a check valve when the C6H6 receiver is on its intake        stroke, and valve Ex3 exhausts the cooled C6H12 from the        regenerator. Note that the exhausts through Ex1 and Ex3 can        be (a) into a slightly higher pressure environment, (b) have        added resistance to opening, or (c) may be mechanically        actuated, ensuring that neither check valve opens prematurely.    -   d. Both receiver mechanisms, having emptied completely, would        now reverse direction, and the cycle would begin again.

Some mixing of remnant C6H12 would occur during the C6H6 thermalregeneration process, and some mixing of remnant C6H6 would occur duringthe C6H12 thermal de-generation process. Also, some amount of H2 wouldbe “mixed in”, since the C6H12+H2 reaction is not likely to be equal to100%. However, because of the relatively small internal area of theregenerator, that mixing can be limited. The net effect on the overallprocess is to reduce the amount of fluid converted per cycle. However,this is expected to only slightly impact the overall thermal efficiencyof the process in a negative way.

In the instance described above, note that pressures for both C6H6 andC6H12 are isobaric and equal at all times. However, the heat exchangeaddition and removal processes may also be partially or whollyisochoric. To create isochoric heating of the C6H6, a displacement willbe required from a third mechanism, or“isochoric displacer”. Note thatthe isochoric displacer would have approximately the same volume andstroke as the mated receiver mechanism, similarly to other displacementprocesses described and shown elsewhere in this document

In another possible STREP variant, the input (and output) of C6H12 canbe isobaric while the input of C6H6 can be isochoric. In that instance,both the pressure and temperature of the C6H6 will increase during theisochoric portion of the displacement process. Note that, in thisinstance, valve Ex2 functions to disallow a high pressure flow back intothe C6H12 receiver mechanism. In addition to Valve Ex2 acting as a checkvalve, regenerator exhaust valve Ex3 will require active sealing againstthe building pressure differential, which indicates that it will need tobe manually operated. Finally, to “match” pressure across Ex2 followingthe two intake processes, Intl will be held open momentarily, allowingpressure to drop in the “displacer” (not shown) as it re-expands trappedC6H6 vapor in the regenerator on the following stroke down to thepressure of the C6H12 receiver mechanism. Note that, with the strokereversal, the C6H12 receiver mechanism will simultaneously begin toincrease pressure, thus “helping” match pressure across valve Ex2. Ex3can also be opened slightly before pressure equalizes between theregenerator and the second receiver mechanism, although that willunavoidably cause a sudden and inefficient pressure drop. A sensor onvalve Ex2 could determine when valve Ex2 begins to move towards open dueto pressure equalization, signaling a solenoid to close In1.

Having a “mixed” isobaric and isochoric process is especially beneficialfor the C6H6+H2 to C6H12 conversion process, since the latent heatrequirement for a given mass of fluid is reduced for an isochoricprocess in relation to an isobaric process. C6H6 is a less dense fluidthan C6H12, and thus requires less latent heat per change in degreetemperature. Using an isochoric flow for the C6H12 delivers less heat.And since the mol count will be ideally equal for both C6H6 and C6H12,such a mixed isobaric and isochoric process is beneficial by allowingthe C6H12 to approach giving up just sufficient latent heat to supplythe requirement for preheating the C6H6 up to the temperature requiredfor the exothermic generation of said C6H12.

In one approach to utilizing this process in a heat engine, the C6H6 isinitially at 1.5 atm and approximately 670 R, while the C6H12 isinitially at 1.5 atm and 1000 R. At the end of the isochoricregeneration process, C6H6 is assumed to be at 1.5 atm and 1000 R andC6H12 is assumed to be at 1.0 atm and 670 R. Note that the latent heatof C6H12 condensation is still available for vaporizing C6H6 liquid.

Of course, the C6H6 vapor thus produced would still require E.R.E.Ccompression, and the required H2 to complete the reaction to C6H112would still require gaseous compression. Since the temperature (1,000 R)at which the endothermic reaction takes place is explicitly tied to thepressure (1.5 atm) at which the endothermic reaction takes place, andsince ideally the only output from the endothermic reactor is C6H12, therequired W-in of pumping, E.R.E.C. compression, and H2 compression makethis process a net consumer of work. That can amount to very little workrequired at close to atmospheric pressure, but those low pressures willalso limit the temperature of the exothermic reaction produced.

Consequently, the work required to produce C6H12 will have to be laidagainst any work produced by the exothermic heat. For example, usingexothermic heat to help power an “isochoric source heating+exothermicpreheating” engine such as has been described above means overallthermal efficiency will be a function of that cycle's W-in requirementplus the W-in requirement of C6H12 production subtracted from the netW-out of the overall process.

Finally, there's the intriguing possibility of a combinedregenerator/exothermic reactor STREP (FIG. 38 ). It's just barelypossible that a catalyst can be integrated with a regenerator. In FIG.38 , the large piston on the left represents preheated vaporous/gaseousexothermic fluid, for example one mol of C6H6 and three moles of H2. Asis clear from U.S. Pat. No. 3,225,538, the conversion to one mol ofC6H12 will occur at both constant pressure and constant temperature.However, volume will be about a quarter the size following theconversion. Hence the much smaller piston on the left that has an intakewhen the larger piston on the left has an exhaust. The single piston onthe right is representing a constant pressure intake and exhaust, takingin a cold high density but completely non-reactive substance through thecatalytic reactor/regenerator and exhausting the hot high densitysubstance on the following stroke.

Because the same pressures are seen for all pistons, the net W-inrequired is only due to friction, pumping, and thermal leakage losses.Ideally, no W-in would be required, outside of the work of pressurizing,and that would be balanced by both W-out and the generation of, andpotential utilization of, thermal energy from the exothermic reactor.

Although specific examples are described herein, the scope of thetechnology is not limited to those specific examples. Moreover, whiledifferent examples and embodiments may be described separately, suchembodiments and examples may be combined with one another inimplementing the technology described herein. One skilled in the artwill recognize other embodiments or improvements that are within thescope and spirit of the present technology. Therefore, the specificexamples disclosed are not to be interpreted in a limiting sense. Thescope of the technology is defined by the following claims andequivalents thereof.

What is claimed is:
 1. For efficiently exchanging heat between twostreams of fluid at approximately equal pressure while simultaneouslyreducing the internal volume and general overall mass of the heatexchange means per quantity of heat exchanged over time, a means termeda Synchronized Thermal Regenerator Exchange Pump (STREP) composed of (1)a piston and cylinder means termed a receiver, said receiver having anintake valve means and an exhaust valve means, (2) a second piston andcylinder means termed a synchronizer, said synchronizer having an intakevalve means and an exhaust valve means, (3) a hollow housing meanscontaining a metallic sponge or regenerator means termed a regenerator,said regenerator having an intake valve means, an exhaust valve means, aport means connected to the exhaust valve of said receiver means, and asecond port means connected to the synchronizer intake valve means, (4)piston movement means such as a crankshaft with a crank throw and aconnecting rod between the crankshaft and a piston connecting pin termeda prime mover, said prime mover able to move said receiver piston andsaid synchronizer piston synchronously such that both pistons will reachTop Dead Center (TDC) and Bottom Dead Center (BDC) simultaneously andcyclically, and (5) a force means for operating said prime mover means,where a charge of fluid at a given temperature is taken in by (a) saidreceiver means through (b) said receiver intake valve means as (c) saidreceiver piston means is moved by said (d) prime mover means from TDC toBDC, and where a second charge of fluid at a different temperature issimultaneously and synchronously taken in by (e) said synchronizerpiston means past (f) the intake valve means of (g) said regeneratormeans, (h) through said regenerator means, (i) out said regenerator portmeans connected to (j) said synchronizer intake valve means, and (k)into said synchronizer means, where in the process of said synchronizercylinder means taking in said charge of fluid heat is either given up tosaid regenerator means or removed from said regenerator means, thusraising or lowering the temperature of the fluid entering saidsynchronizer means, and where, upon reaching BDC, both pistonssimultaneously and synchronously (l) reverse direction and begin movingtowards TDC by action of said prime mover means, whereby said receivercylinder, synchronizer cylinder, and regenerator means' intake valvemeans ideally instantaneously, simultaneously and synchronously (m)close as said receiver cylinder, synchronizer cylinder, and regeneratormeans' exhaust valve means ideally instantaneously, simultaneously andsynchronously (n) open, resulting in the simultaneous (o) expulsion offluid out of said receiver and synchronizer cylinder means by (p) actionof said piston means driven by (q) said prime mover means, in the caseof said synchronizer cylinder means its fluid being (r) driven out ofits exhaust valve means while in the case of said receiver cylindermeans its fluid being (s) driven out of its exhaust valve means, (t)through said regenerator means, and (u) out said regenerator exhaustvalve means, where in the process of said receiver means (v) passingfluid through said regenerator means its charge of fluid will either (w)receive heat from or give up heat to said regenerator means, said heathaving been earlier either (x) deposited by or removed from saidregenerator means but in any case the opposite effect achieved by saidsynchronizer means, thus resulting in an efficient exchange of heatbetween the two streams at approximately equal pressure whilesimultaneously reducing the internal volume per quantity of heatexchanged over time and general overall mass.